Varying Compression Ratios in Energy Storage and Retrieval Systems

ABSTRACT

The present disclosure provides pumped thermal energy storage systems that can be used to store electrical energy. A pumped thermal energy storage system of the present disclosure can store energy by operating as a heat pump or refrigerator, whereby net work input can be used to transfer heat from the cold side to the hot side. A working fluid of the system is capable of efficient heat exchange with heat storage fluids on a hot side of the system and on a cold side of the system. The system can extract energy by operating as a heat engine transferring heat from the hot side to the cold side, which can result in net work output. Systems of the present disclosure can employ solar heating for improved storage efficiency.

CROSS-REFERENCE

This application is a continuation U.S. patent application Ser. No.14/668,610, filed Mar. 25, 2015 (“SYSTEMS AND METHODS FOR ENERGY STORAGEAND RETRIEVAL”), which claims priority to Patent Cooperation TreatyApplication Serial No. PCT/US2013/062469 filed Sep. 27, 2013 (“SYSTEMSAND METHODS FOR ENERGY STORAGE AND RETRIEVAL”), which claims priority toU.S. Provisional Patent Application Ser. No. 61/706,337, filed Sep. 27,2012 (“SYSTEMS AND METHODS FOR ADIABATIC SALT ENERGY STORAGE”), and U.S.Provisional Patent Application Ser. No. 61/868,070, filed Aug. 20, 2013(“SYSTEMS AND METHODS FOR ENERGY STORAGE AND RETRIEVAL”), each of whichis entirely incorporated herein by reference.

BACKGROUND

An energy storage system is capable of storing one or more forms ofenergy for later extraction and use. In some instances, energy storagesystems can employ one or more energy conversion processes in order tostore and extract energy. Energy storage systems can be charged(storage) and discharged (extraction). Some energy storage systems areused to store electrical energy.

Energy storage systems may be employed for use in delivering electricalenergy to or removing electrical energy from an electrical energydistribution system, such as a power grid. A power grid may be used todeliver electrical energy to end users.

SUMMARY

This disclosure provides pumped thermal energy storage systems (also“energy storage systems,” “pumped thermal systems”, “pumped thermalelectric systems” or “systems” herein). Energy inputs to the systemsprovided herein can include electrical energy (electricity), in somecases converted from non-electrical types of energy (such aselectromagnetic energy (radiation) or mechanical energy), thermal energy(heat), chemical energy, mechanical energy, in some cases converted fromelectrical energy, or a combination thereof. Energy outputs from thesystems provided herein can include electrical energy (electricity)converted from mechanical energy, thermal energy (heat), mechanicalenergy, or a combination thereof. For instance, electrical energy can bestored by the systems of the disclosure. In some cases, the electricalenergy can be input in combination with thermal energy provided throughsolar heating or combustion of a combustible substance such as, forexample, a fossil fuel or biomass. In some situations, the systemsherein can be augmented by or be directly utilized as energy conversionsystems through the use of thermal energy from combusting a fossil fuelor biomass.

Large scale energy storage systems can be utilized in power generationand distribution systems. Existing alternatives include pumpedhydroelectric storage for storing electrical energy, and solar thermalsystems for storing thermal energy (as sensible energy).

Recognized herein is the need to provide large scale energy storagesystems with high efficiency and reliability, and at low capital costs.The energy storage systems of the disclosure are not geographically orresource-limited and are capable of storing electrical energy as wellas, in some cases, electromagnetic energy or thermal energy (as sensibleenergy). Furthermore, the systems of the disclosure can be augmented byadditional energy conversion processes, or be directly utilized asenergy conversion systems (without energy storage). In some instances,the systems of the disclosure may accept or provide one or more wasteenergy streams, such as waste heat or waste cold streams.

Energy storage systems of the disclosure can include a working fluidflowing in a closed cycle. In some cases, the closed cycle can include acompressor and a turbine. The working fluid can be capable of efficientheat exchange with heat storage fluids on a hot side of the system andon a cold side of the system. The system can operate as a heat engine bytransferring heat from the hot side to the cold side, resulting in netmechanical work output. The system can also operate as a heat pump orrefrigerator, whereby net mechanical work input is used to transfer heatfrom the cold side to the hot side. The mechanical work inputs and/oroutputs may be converted from/to electrical work using amotor/generator. The compressor, the turbine and the motor/generator canbe located on a common shaft. Heat exchange (sensible energy transfer)between the working fluid of the system and the heat storage fluids canoccur in counter-flow heat exchangers. The hot side and cold side heatstorage fluids can each have a corresponding pair of storage tanks,where heat transfer to/from a heat storage fluid entails flow of theheat storage fluid between its two corresponding storage tanks. Whiletraveling from a first storage tank to a second tank, the heat storagefluid can enter a heat exchanger where it either receives or donatesheat to the working fluid of a thermodynamic cycle. Each heat storagefluid with corresponding heat exchanger and pair of storage tanks canconstitute a heat storage unit.

An aspect of the present disclosure provides an energy storage andretrieval system comprising a compressor, a first heat storage unit, aturbine, a second heat storage unit, a working fluid that flows along afluid flow path in a closed cycle including, in sequence, thecompressor, the first heat storage unit, the turbine, and the secondheat storage unit, and an auxiliary tank comprising the working fluid.The auxiliary tank is in fluid communication with the fluid flow path ofthe closed cycle. The system alternately operates as both (i) a heatengine to provide mechanical work from heat and (ii) as a heat pump touse mechanical work to store heat.

Another aspect of the present disclosure provides a method for storingand releasing energy comprising (a) providing a system comprising aclosed cycle comprising, in sequence, a compressor, a first heat storageunit, a turbine, and a second heat storage unit. The system furthercomprises an auxiliary tank for a working fluid. The first and secondheat storage units exchange heat with the working fluid flowing throughthe closed cycle. The method further comprises (b) alternately andsequentially operating the system in a refrigerator mode and a heatengine mode. In the refrigerator mode, mechanical work is used totransfer thermal energy from the second heat storage unit to the firstheat storage unit. In the heat engine mode, thermal energy transferredfrom the first heat storage unit to the second heat storage unit is usedto provide mechanical work.

Another aspect of the present disclosure is directed to a method. forstoring and releasing energy comprising (a) increasing the pressure of aworking fluid, operating in a closed cycle, from a first pressure to asecond pressure with the aid of a compressor, thereby increasing thetemperature of the working fluid. The method further comprises (b) usinga first heat storage unit downstream of the compressor and in thermalcommunication with the working fluid for (i) in a storing mode, removingheat from the working fluid and decreasing the temperature of theworking fluid, which decrease in temperature is at substantially thesecond pressure, or (ii) in a releasing mode, supplying heat to theworking fluid and increasing the temperature of the working fluid, whichincrease in temperature is at substantially the second pressure. Themethod further comprises (c) decreasing the pressure of the workingfluid from the second pressure to the first pressure with the aid of aturbine, thereby decreasing the temperature of the working fluid. Themethod further comprises (d) using a second heat storage unit downstreamof the turbine and in thermal communication with the working fluid for(i) in a storing mode, supplying heat to the working fluid andincreasing the temperature of the working fluid, which decrease intemperature is at substantially the first pressure, or (ii) in areleasing mode, removing heat from the working fluid and decreasing thetemperature of the working fluid, which decrease in temperature is atsubstantially the first pressure. The working fluid does not undergo aphase change. Heat is supplied to or removed from the working fluid withthe aid of a recuperator.

A further aspect of the present disclosure provides a pumped thermalenergy storage system comprising a closed fluid flow path configured tocirculate a working fluid. A first thermal storage medium is in thermalcommunication with the working fluid. The first thermal storage mediumis capable of exchanging heat with the working fluid between a first lowtemperature and a first high temperature. A second thermal storagemedium is in thermal communication with the working fluid. The secondthermal store medium is capable of exchanging heat with the workingfluid between a second low temperature and a second high temperature.The first low temperature is higher than the second high temperature. Arecuperator is in fluid communication with the fluid flow path. Therecuperator exchanges heat with the working fluid between the first lowtemperature and the second high temperature. The working fluid does notundergo a phase change.

A further aspect of the present disclosure is directed to a system forstoring and extracting electrical energy comprising an electrical energyinput and an electrical energy output. A ratio of the electrical energyoutput to the electrical energy input is greater than 1. The systemstores or extracts electrical energy using at least one heat exchangersystem comprising a thermal storage fluid that directs thermal energyinto or extracts thermal energy from a circulatory fluid flow path. Thesystem comprises a controller that is programmed to regulate (i) atemperature difference between any two thermally coupled fluid elementsof the heat exchanger system, and/or (ii) one or more fluid propertiesof the fluid elements of the heat exchanger system such that entropygeneration in the heat exchanger system is minimized during operation ofthe system.

Another aspect of the present disclosure is directed to a method forstoring electrical energy comprising alternately and sequentially (i)providing an electrical work input to a heat pump in thermalcommunication with two thermal storage media, thereby storing energy,and (ii) providing an electrical work output from a heat engine inthermal communication with the two thermal storage media, therebyextracting energy. The heat pump and the heat engine comprise a workingfluid flowing through a circulatory fluid flow path. The electricalinput or the electrical output is varied by varying an absolute pressureof the working fluid.

In another aspect of the present disclosure, an energy storage andconversion system is provided. The energy storage and conversion systemcomprises a compressor, a first heat exchanger, a turbine, a second heatexchanger, a working fluid that flows along a fluid flow path in aclosed cycle including, in sequence, the compressor, the first heatexchanger unit, the turbine, and the second heat exchanger, and a thirdheat exchanger in thermal communication with the working fluid. Thesystem alternately operates as both (i) an electricity storage andretrieval cycle, and (ii) a heat engine to ambient cycle. In theelectricity storage and retrieval cycle, the first heat exchangerexchanges heat between the working fluid and a first thermal storagemedium and the second heat exchanger exchanges heat between the workingfluid and a second thermal storage medium. In the heat engine to ambientcycle, the first heat exchanger transfers heat from the first thermalstorage medium to the working fluid and the third heat exchangertransfers heat from the working fluid to the ambient environment.

Another aspect of the present disclosure relates to an energy storageand conversion system comprising a compressor a first heat exchanger, aturbine, a second heat exchanger, a working fluid that flows along afluid flow path in a closed cycle including, in sequence, thecompressor, the first heat exchanger unit, the turbine, and the secondheat exchanger, and a third heat exchanger in thermal communication withthe working fluid. The system alternately operates as both (i) anelectricity storage and retrieval cycle, and (ii) a refrigerator toambient cycle. In the electricity storage and retrieval cycle, the firstheat exchanger exchanges heat between the working fluid and a firstthermal storage medium and the second heat exchanger exchanges heatbetween the working fluid and a second thermal storage medium. In therefrigerator to ambient cycle, the third heat exchanger transfers heatfrom the working fluid to the ambient environment and the second heatexchanger transfers heat from the second thermal storage medium to theworking fluid.

Additional aspects and advantages of the present disclosure will becomereadily apparent to those skilled in this art from the followingdetailed description, wherein only illustrative embodiments of thepresent disclosure are shown and described. As will be realized, thepresent disclosure is capable of other and different embodiments, andits several details are capable of modifications in various obviousrespects, all without departing from the disclosure. Accordingly, thedrawings and description are to be regarded as illustrative in nature,and not as restrictive.

INCORPORATION BY REFERENCE

All publications, patents, and patent applications mentioned in thisspecification are herein incorporated by reference to the same extent asif each individual publication, patent, or patent application wasspecifically and individually indicated to be incorporated by reference.

BRIEF DESCRIPTION OF DRAWINGS

The novel features of the invention are set forth with particularity inthe appended claims. A better understanding of the features andadvantages of the present invention will be obtained by reference to thefollowing detailed description that sets forth illustrative embodiments,in which the principles of the invention are utilized, and theaccompanying drawings or figures (also “FIG.” and “FIGs.” herein), ofwhich:

FIG. 1 schematically illustrates operation of a pumped thermal electricstorage system.

FIG. 2A is a schematic flow diagram of working fluid and heat storagemedia of a pumped thermal system in a charge/heat pump mode.

FIG. 2B is a schematic flow diagram of working fluid and heat storagemedia of a pumped thermal system in a discharge/heat engine mode.

FIG. 3A is a schematic pressure and temperature diagram of the workingfluid as it undergoes the charge cycle in FIG. 2A.

FIG. 3B is a schematic pressure and temperature diagram of the workingfluid as it undergoes the discharge cycle in FIG. 2B.

FIG. 4A is a schematic perspective view of a closed working fluid systemin the pumped thermal system in FIGS. 2A-2B.

FIG. 4B is a schematic perspective view of the pumped thermal system inFIGS. 2A-2B with hot side and cold side storage tanks and a closed cycleworking fluid system.

FIG. 5A shows a heat storage charge cycle for a water/molten salt systemwith η_(c)=0.9 and η_(t)=0.95. The dashed lines correspond to η_(c)=1.

FIG. 5B shows a heat storage discharge (extraction) cycle for thewater/molten salt system in FIG. 5A with η_(c)=0.9 and η_(t)=0.95. Thedashed lines correspond to η_(c)=η_(t)=1.

FIG. 5C shows a heat storage cycle in a pumped thermal system withvariable compression ratios between the charge and discharge cycles.

FIG. 6A shows roundtrip efficiency contours for a water/salt system. Thesymbols ⊕ and

represent an approximate range of present large turbomachinery adiabaticefficiency values. The dashed arrows represent the direction ofincreasing efficiency.

FIG. 6B shows roundtrip efficiency contours for a colder storage/saltsystem. The symbols ⊕

and represent an approximate range of present large turbomachineryadiabatic efficiency values.

FIG. 7A is a schematic flow diagram of working fluid and heat storagemedia of a pumped thermal system with a gas-gas heat exchanger for theworking fluid in a charge/heat pump mode.

FIG. 7B is a schematic flow diagram of working fluid and heat storagemedia of a pumped thermal system with a gas-gas heat exchanger for theworking fluid in a discharge/heat engine mode.

FIG. 7C is a schematic flow diagram of working fluid and heat storagemedia of a pumped thermal system with a gas-gas heat exchanger for theworking fluid in a charge/heat pump mode with indirect heat rejection tothe environment.

FIG. 7D is a schematic flow diagram of working fluid and heat storagemedia of a pumped thermal system with a gas-gas heat exchanger for theworking fluid in a discharge/heat engine mode with indirect heatrejection to the environment.

FIG. 7E is a schematic diagram of a piping and valve configuration forthe working fluid in the counter-flow thermodynamic charge cycle of FIG.7C. The circular black and white symbols represent three-way or four-wayvalves. Arrows in pipes represent working fluid flow.

FIG. 7F is a schematic diagram of a piping and valve configuration forthe working fluid in the counter-flow thermodynamic discharge cycle ofFIG. 7D. The circular black and white symbols represent three-way orfour-way valves. Arrows in pipes represent working fluid flow.

FIG. 8A shows a heat storage charge cycle for a storage system with agas-gas heat exchanger, a cold side storage medium capable of going downto temperatures significantly below ambient temperature and η_(c)=0.9and η_(t)=0.95.

FIG. 8B shows a heat storage discharge cycle for a storage system with agas-gas heat exchanger, a cold side storage medium capable of going downto temperatures significantly below ambient temperature and η_(c)=0.9and η_(t)=0.95.

FIG. 9A is a schematic flow diagram of cold side recharging in a pumpedheat cycle.

FIG. 9B is a schematic flow diagram of cold side recharging in a pumpedheat cycle with a heat exchanger to ambient.

FIG. 9C is a schematic flow diagram of hot side recharging in a pumpedheat cycle in solar mode with heating of a solar salt solely by solarpower.

FIG. 9D is a schematic flow diagram of hot side recharging in a pumpedheat cycle in solar mode using a heat exchanger between an intermediatefluid tank and a solar salt.

FIG. 9E is a schematic flow diagram of a pumped thermal system chargewith a gas-gas heat exchanger in parallel with solar heat input.

FIG. 9F is a schematic flow diagram of cold side recharging in a pumpedheat cycle with hot side heat rejection to an intermediate fluidcirculated in a thermal bath.

FIG. 9G is a schematic flow diagram of a pumped thermal system dischargecycle with heat rejection to ambient.

FIG. 9H is a schematic flow diagram of a pumped thermal system dischargecycle with heat rejection to an intermediate fluid circulated in athermal bath at ambient temperature.

FIG. 10 shows a cold side recharge cycle for a hexane/salt system insolar mode in accordance with a modified system of FIG. 9A withη_(c)=0.9 and η_(t)=0.95.

FIGS. 11A-D show (clockwise) cold side recharging in a pumped heat cycleusing intercooling with 0, 1, 2 and 3 stages.

FIG. 12 illustrates the effect of additional stages of intercooling onsolar charge cycle compression work for a system with r=3.38 and T₁=300K (27° C.). The isothermal compression limit is represented by ahorizontal line.

FIG. 13A is a schematic diagram of an exemplary piping and valveconfiguration for achieving the electricity only charge cycle in FIG.7C. The circular black and white symbols represent three-way or four-wayvalves. Arrows in pipes represent working fluid flow.

FIG. 13B is a schematic diagram of an exemplary piping and valveconfiguration for achieving the electricity only discharge cycle in FIG.7D. The circular black and white symbols represent three-way or four-wayvalves. Arrows in pipes represent working fluid flow.

FIG. 13C is a schematic diagram of an exemplary piping and valveconfiguration for achieving the cold side recharging refrigeration cyclein FIG. 9F or 9A. The circular black and white symbols representthree-way or four-way valves. Arrows in pipes represent working fluidflow.

FIG. 13D is a schematic diagram of an exemplary piping and valveconfiguration for achieving heat engine to ambient cycle in FIG. 9H or9G with addition of a gas-gas heat exchanger. The circular black andwhite symbols represent three-way or four-way valves. Arrows in pipesrepresent working fluid flow.

FIGS. 14A and 14B are pumped thermal systems with separatecompressor/turbine pairs for charge and discharge modes.

FIGS. 15A and 15B show pumped thermal systems configured in a combustionheat input generation mode.

FIG. 15C is a schematic flow diagram of hot side recharging in a pumpedheat cycle through heating by a combustion heat source or a waste heatsource.

FIG. 16 shows an example of a pumped thermal system with pressureregulated power control.

FIG. 17 shows an example of a pumped thermal system with a pressureencased generator.

FIG. 18 is an example of variable stators in a compressor/turbine pair.

FIG. 19 shows a computer system that is programmed to implement variousmethods and/or regulate various systems of the present disclosure.

DETAILED DESCRIPTION

While various embodiments of the invention have been shown and describedherein, it will be obvious to those skilled in the art that suchembodiments are provided by way of example only. Numerous variations,changes, and substitutions may occur to those skilled in the art withoutdeparting from the invention. It should be understood that variousalternatives to the embodiments of the invention described herein may beemployed. It shall be understood that different aspects of the inventioncan be appreciated individually, collectively, or in combination witheach other.

The term “reversible,” as used herein, generally refers to a process oroperation that can be reversed via infinitesimal changes in someproperty of the process or operation without substantial entropyproduction (e.g., dissipation of energy). A reversible process may beapproximated by a process that is at thermodynamic equilibrium. In someexamples, in a reversible process, the direction of flow of energy isreversible. As an alternative, or in addition to, the general directionof operation of a reversible process (e.g., the direction of fluid flow)can be reversed, such as, e.g., from clockwise to counterclockwise, andvice versa.

The term “sequence,” as used herein, generally refers to elements (e.g.,unit operations) in order. Such order can refer to process order, suchas, for example, the order in which a fluid flows from one element toanother. In an example, a compressor, heat storage unit and turbine insequence includes the compressor upstream of the heat exchange unit, andthe heat exchange unit upstream of the turbine. In such a case, a fluidcan flow from the compressor to the heat exchange unit and from the heatexchange unit to the turbine. A fluid flowing through unit operations insequence can flow through the unit operations sequentially. A sequenceof elements can include one or more intervening elements. For example, asystem comprising a compressor, heat storage unit and turbine insequence can include an auxiliary tank between the compressor and theheat storage unit. A sequence of elements can be cyclical.

Pumped Thermal Systems

The disclosure provides pumped thermal systems capable of storingelectrical energy and/or heat, and releasing energy (e.g., producingelectricity) at a later time. The pumped thermal systems of thedisclosure may include a heat engine, and a heat pump (or refrigerator).In some cases, the heat engine can be operated in reverse as a heatpump. In some cases, the heat engine can be operated in reverse as arefrigerator. Any description of heat pump/heat engine systems orrefrigerator/heat engine systems capable of reverse operation herein mayalso be applied to systems comprising separate and/or a combination ofseparate and reverse-operable heat engine system(s), heat pump system(s)and/or refrigerator system(s). Further, as heat pumps and refrigeratorsshare the same operating principles (albeit with differing objectives),any description of configurations or operation of heat pumps herein mayalso be applied to configurations or operation of refrigerators, andvice versa.

Systems of the present disclosure can operate as heat engines or heatpumps (or refrigerators). In some situations, systems of the disclosurecan alternately operate as heat engines and heat pumps. In someexamples, a system can operate as a heat engine to generate power, andsubsequently operate as a heat pump to store energy, or vice versa. Suchsystems can alternately and sequentially operate as heat engines as heatpumps. In some cases, such systems reversibly or substantiallyreversibly operate as heat engines as heat pumps.

Reference will now be made to the figures, wherein like numerals referto like parts throughout. It will be appreciated that the figures andfeatures therein are not necessarily drawn to scale.

FIG. 1 schematically illustrates operating principles of pumped thermalelectric storage using a heat pump/heat engine electricity storagesystem. Electricity may be stored in the form of thermal energy of twomaterials or media at different temperatures (e.g., thermal energyreservoirs comprising heat storage fluids or thermal storage media) byusing a combined heat pump/heat engine system. In a charging or heatpump mode, work may be consumed by the system for transferring heat froma cold material or medium to a hot material or medium, thus lowering thetemperature (e.g., sensible energy) of the cold material and increasingthe temperature (i.e., sensible energy) of the hot material. In adischarging or heat engine mode, work may be produced by the system bytransferring heat from the hot material to the cold material, thuslowering the temperature (i.e., sensible energy) of the hot material andincreasing the temperature (i.e., sensible energy) of the cold material.The system may be configured to ensure that the work produced by thesystem on discharge is a favorable fraction of the energy consumed oncharge. The system may be configured to achieve high roundtripefficiency, defined herein as the work produced by the system ondischarge divided by the work consumed by the system on charge. Further,the system may be configured to achieve the high roundtrip efficiencyusing components of a desired (e.g., acceptably low) cost. Arrows H andW in FIG. 1 represent directions of heat flow and work, respectively.

Heat engines, heat pumps and refrigerators of the disclosure may involvea working fluid to and from which heat is transferred while undergoing athermodynamic cycle. The heat engines, heat pumps and refrigerators ofthe disclosure may operate in a closed cycle. Closed cycles allow, forexample, a broader selection of working fluids, operation at elevatedcold side pressures, operation at lower cold side temperatures, improvedefficiency, and reduced risk of turbine damage. One or more aspects ofthe disclosure described in relation to systems having working fluidsundergoing closed cycles may also be applied to systems having workingfluids undergoing open cycles.

In one example, the heat engines may operate on a Brayton cycle and theheat pumps/refrigerators may operate on a reverse Brayton cycle (alsoknown as a gas refrigeration cycle). Other examples of thermodynamiccycles that the working fluid may undergo or approximate include theRankine cycle, the ideal vapor-compression refrigeration cycle, theStirling cycle, the Ericsson cycle or any other cycle advantageouslyemployed in concert with heat exchange with heat storage fluids of thedisclosure.

The working fluid can undergo a thermodynamic cycle operating at one,two or more pressure levels. For example, the working fluid may operatein a closed cycle between a low pressure limit on a cold side of thesystem and a high pressure limit on a hot side of the system. In someimplementations, a low pressure limit of about 10 atmospheres (atm) orgreater can be used. In some instances, the low pressure limit may be atleast about 1 atm, at least about 2 atm, at least about 5 atm, at leastabout 10 atm, at least about 15 atm, at least about 20 atm, at leastabout 30 atm, at least about 40 atm, at least about 60 atm, at leastabout 80 atm, at least about 100 atm, at least about 120 atm, at leastabout 160 atm, or at least about 200 atm, 500 atm, 1000 atm, or more. Insome instances, a sub-atmospheric low pressure limit may be used. Forexample, the low pressure limit may be less than about 0.1 atm, lessthan about 0.2 atm, less than about 0.5 atm, or less than about 1 atm.In some instances, the low pressure limit may be about 1 atmosphere(atm). In the case of a working fluid operating in an open cycle, thelow pressure limit may be about 1 atm or equal to ambient pressure.

In some cases, the value of the low pressure limit may be selected basedon desired power output and/or power input requirements of thethermodynamic cycle. For example, a pumped thermal system with a lowpressure limit of about 10 atm may be able to provide a power outputcomparable to an industrial gas turbine with ambient (1 atm) air intake.The value of the low pressure limit may also be subject to cost/safetytradeoffs. Further, the value of the low pressure limit may be limitedby the value of the high pressure limit, the operating ranges of the hotside and cold side heat storage media (e.g., pressure and temperatureranges over which the heat storage media are stable), pressure ratiosand operating conditions (e.g., operating limits, optimal operatingconditions, pressure drop) achievable by turbomachinery and/or othersystem components, or any combination thereof. The high pressure limitmay be determined in accordance with these system constraints. In someinstances, higher values of the high pressure limit may lead to improvedheat transfer between the working fluid and the hot side storage medium.

Working fluids used in pumped thermal systems may include air, argon,other noble gases, carbon dioxide, hydrogen, oxygen, or any combinationthereof, and/or other fluids in gaseous, liquid, critical, orsupercritical state (e.g., supercritical CO₂). The working fluid can bea gas or a low viscosity liquid (e.g., viscosity below about 500×10⁻⁶Poise at 1 atm), satisfying the requirement that the flow be continual.In some implementations, a gas with a high specific heat ratio may beused to achieve higher cycle efficiency than a gas with a low specificheat ratio. For example, argon (e.g., specific heat ratio of about 1.66)may be used to substitute air (e.g., specific heat ratio of about 1.4).In some cases, the working fluid may be a blend of one, two, three ormore fluids. In one example, helium (having a high thermal conductivityand a high specific heat) may be added to the working fluid (e.g.,argon) to improve heat transfer rates in heat exchangers.

Pumped thermal systems herein may utilize heat storage media ormaterials, such as one or more heat storage fluids. The heat storagemedia can be gases or low viscosity liquids, satisfying the requirementthat the flow be continual. The systems may utilize a first heat storagemedium on a hot side of the system (“hot side thermal storage (HTS)medium” or “HTS” herein) and a second heat storage medium on a cold sideof the system (“cold side thermal storage (CTS) medium” or “CTS”herein). The thermal storage media (e.g., low viscosity liquids) canhave high heat capacities per unit volume (e.g., heat capacities aboveabout 1400 Joule (kilogram Kelvin)⁻¹) and high thermal conductivities(e.g., thermal conductivities above about 0.7 Watt (meter Kelvin)⁻¹). Insome implementations, several different thermal storage media (also“heat storage media” herein) on either the hot side, cold side or boththe hot side and the cold side may be used.

The operating temperatures of the hot side thermal storage medium can bein the liquid range of the hot side thermal storage medium, and theoperating temperatures of the cold side thermal storage medium can be inthe liquid range of the cold side thermal storage medium. In someexamples, liquids may enable a more rapid exchange of large amounts ofheat by convective counter-flow than solids or gases. Thus, in somecases, liquid HTS and CTS media may advantageously be used. Pumpedthermal systems utilizing thermal storage media herein mayadvantageously provide a safe, non-toxic and geography-independentenergy (e.g., electricity) storage alternative.

In some implementations, the hot side thermal storage medium can be amolten salt or a mixture of molten salts. Any salt or salt mixture thatis liquid over the operating temperature range of the hot side thermalstorage medium may be employed. Molten salts can provide numerousadvantages as thermal energy storage media, such as low vapor pressure,lack of toxicity, chemical stability, low chemical reactivity withtypical steels (e.g., melting point below the creep temperature ofsteels, low corrosiveness, low capacity to dissolve iron and nickel),and low cost. In one example, the HTS is a mixture of sodium nitrate andpotassium nitrate. In some examples, the HTS is a eutectic mixture ofsodium nitrate and potassium nitrate. In some examples, the HTS is amixture of sodium nitrate and potassium nitrate having a lowered meltingpoint than the individual constituents, an increased boiling point thanthe individual constituents, or a combination thereof. Other examplesinclude potassium nitrate, calcium nitrate, sodium nitrate, sodiumnitrite, lithium nitrate, mineral oil, or any combination thereof.Further examples include any gaseous (including compressed gases),liquid or solid media (e.g., powdered solids) having suitable (e.g.,high) thermal storage capacities and/or capable of achieving suitable(e.g., high) heat transfer rates with the working fluid. For example, amix of 60% sodium nitrate and 40% potassium nitrate (also referred to asa solar salt in some situations) can have a heat capacity ofapproximately 1500 Joule (Kelvin mole)⁻¹ and a thermal conductivity ofapproximately 0.75 Watt (meter Kelvin)⁻¹ within a temperature range ofinterest. The hot side thermal storage medium may be operated in atemperature range that structural steels can handle.

In some cases, liquid water at temperatures of about 0° C. to 100° C.(about 273 K-373 K) and a pressure of about 1 atm may be used as thecold side thermal storage medium. Due to a possible explosion hazardassociated with presence of steam at or near the boiling point of water,the operating temperature can be kept below about 100° C. or less whilemaintaining an operating pressure of 1 atm (i.e., no pressurization). Insome cases, the temperature operating range of the cold side thermalstorage medium may be extended (e.g., to −30° C. to 100° C. at 1 atm) byusing a mixture of water and one or more antifreeze compounds (e.g.,ethylene glycol, propylene glycol, or glycerol).

As described in greater detail elsewhere herein, improved storageefficiency may be achieved by increasing the temperature difference atwhich the system operates, for example, by using a cold side heatstorage fluid capable of operating at lower temperatures. In someexamples, the cold side thermal storage media may comprise hydrocarbons,such as, for example, alkanes (e.g., hexane or heptane), alkenes,alkynes, aldehydes, ketones, carboxylic acids (e.g., HCOOH), ethers,cycloalkanes, aromatic hydrocarbons, alcohols (e.g., butanol), othertype(s) of hydrocarbon molecules, or any combinations thereof. In somecases, the cold side thermal storage medium can be hexane (e.g.,n-hexane). Hexane has a wide liquid range and can remain fluid (i.e.,runny) over its entire liquid range (−94° C. to 68° C. at 1 atm).Hexane's low temperature properties are aided by its immiscibility withwater. Other liquids, such as, for example, ethanol or methanol canbecome viscous at the low temperature ends of their liquid ranges due topre-crystallization of water absorbed from air. In some cases, the coldside thermal storage medium can be heptane (e.g., n-heptane). Heptanehas a wide liquid range and can remain fluid (i.e., runny) over itsentire liquid range (−91° C. to 98° C. at 1 atm). Heptane's lowtemperature properties are aided by its immiscibility with water. Ateven lower temperatures, other heat storage media can be used, such as,for example, isohexane (2-methylpentane). In some examples, cryogenicliquids having boiling points below about −150° C. (123 K) or about−180° C. (93.15 K) may be used as cold side thermal storage media (e.g.,propane, butane, pentane, nitrogen, helium, neon, argon and krypton,air, hydrogen, methane, or liquefied natural gas). In someimplementations, choice of cold side thermal storage medium may belimited by the choice of working fluid. For example, when a gaseousworking fluid is used, a liquid cold side thermal storage medium havinga liquid temperature range at least partially or substantially above theboiling point of the working fluid may be required.

In some cases, the operating temperature range of CTS and/or HTS mediacan be changed by pressurizing (i.e., raising the pressure) orevacuating (i.e., lowering the pressure) the tanks and thus changing thetemperature at which the storage media undergo phase transitions (e.g.,going from liquid to solid, or from liquid to gas).

In some cases, the hot side and the cold side heat storage fluids of thepumped thermal systems are in a liquid state over at least a portion ofthe operating temperature range of the energy storage device. The hotside heat storage fluid may be liquid within a given range oftemperatures. Similarly, the cold side heat storage fluid may be liquidwithin a given range of temperatures. The heat storage fluids may beheated, cooled or maintained to achieve a suitable operating temperatureprior to, during or after operation.

Pumped thermal systems of the disclosure may cycle between charged anddischarged modes. In some examples, the pumped thermal systems can befully charged, partially charged or partially discharged, or fullydischarged. In some cases, cold side heat storage may be charged (also“recharged” herein) independently from hot side heat storage. Further,in some implementations, charging (or some portion thereof) anddischarging (or some portion thereof) can occur simultaneously. Forexample, a first portion of a hot side heat storage may be rechargedwhile a second portion of the hot side heat storage together with a coldside heat storage are being discharged.

The pumped thermal systems may be capable of storing energy for a givenamount of time. In some cases, a given amount of energy may be storedfor at least about 1 second, at least about 30 seconds, at least about 1minute, at least about 5 minutes, at least about 30 minutes, at leastabout 1 hour, at least about 2 hours, at least about 3 hours, at leastabout 4 hours, at least about 5 hours, at least about 6 hours, at leastabout 7 hours, at least about 8 hours, at least about 9 hours, at leastabout 10 hours, at least about 12 hours at least about 14 hours, atleast about 16 hours, at least about 18 hours, at least about 20 hours,at least about 22 hours, at least about 24 hours (1 day), at least about2 days, at least about 4 days, at least about 6 days, at least about 8days, at least about 10 days, 20 days, 30 days, 60 days, 100 days, 1year or more.

Pumped thermal systems of the disclosure may be capable ofstoring/receiving input of, and/or extracting/providing output of asubstantially large amount of energy and/or power for use with powergeneration systems (e.g., intermittent power generation systems such aswind power or solar power), power distribution systems (e.g. electricalgrid), and/or other loads or uses in grid-scale or stand-alone settings.During a charge mode of a pumped thermal system, electric power receivedfrom an external power source (e.g., a wind power system, a solarphotovoltaic power system, an electrical grid etc.) can be used operatethe pumped thermal system in a heat pump mode (i.e., transferring heatfrom a low temperature reservoir to a high temperature reservoir, thusstoring energy). During a discharge mode of the pumped thermal system,the system can supply electric power to an external power system or load(e.g., one or more electrical grids connected to one or more loads, aload, such as a factory or a power-intensive process, etc.) by operatingin a heat engine mode (i.e., transferring heat from a high temperaturereservoir to a low temperature reservoir, thus extracting energy). Asdescribed elsewhere herein, during charge and/or discharge, the systemmay receive or reject thermal power, including, but not limited toelectromagnetic power (e.g., solar radiation) and thermal power (e.g.,sensible energy from a medium heated by solar radiation, heat ofcombustion etc.).

In some implementations, the pumped thermal systems aregrid-synchronous. Synchronization can be achieved by matching speed andfrequency of motors/generators and/or turbomachinery of a system withthe frequency of one or more grid networks with which the systemexchanges power. For example, a compressor and a turbine can rotate at agiven, fixed speed (e.g., 3600 revolutions per minute (rpm)) that is amultiple of grid frequency (e.g., 60 hertz (Hz)). In some cases, such aconfiguration may eliminate the need for additional power electronics.In some implementations, the turbomachinery and/or the motors/generatorsare not grid synchronous. In such cases, frequency matching can beaccomplished through the use of power electronics. In someimplementations, the turbomachinery and/or the motors/generators are notdirectly grid synchronous but can be matched through the use of gearsand/or a mechanical gearbox. As described in greater detail elsewhereherein, the pumped thermal systems may also be rampable. Suchcapabilities may enable these grid-scale energy storage systems tooperate as peaking power plants and/or as a load following power plants.In some cases, the systems of the disclosure may be capable of operatingas base load power plants.

Pumped thermal systems can have a given power capacity. In some cases,power capacity during charge may differ from power capacity duringdischarge. For example, each system can have a charge and/or dischargepower capacity of less than about 1 megawatt (MW), at least about 1megawatt, at least about 2 MW, at least about 3 MW, at least about 4 MW,at least about 5 MW, at least about 6 MW, at least about 7 MW, at leastabout 8 MW, at least about 9 MW, at least about 10 MW, at least about 20MW, at least about 30 MW, at least about 40 MW, at least about 50 MW, atleast about 75 MW, at least about 100 MW, at least about 200 MW, atleast about 500 MW, at least about 1 gigawatt (GW), at least about 2 GW,at least about 5 GW, at least about 10 GW, at least about 20 GW, atleast about 30 GW, at least about 40 GW, at least about 50 GW, at leastabout 75 GW, at least about 100 GW, or more.

Pumped thermal systems can have a given energy storage capacity. In oneexample, a pumped thermal system is configured as a 100 MW unitoperating for 10 hours. In another example, a pumped thermal system isconfigured as a 1 GW plant operating for 12 hours. In some instances,the energy storage capacity can be less than about 1 megawatt hour(MWh), at least about 1 megawatt hour, at least about 10 MWh, at leastabout 100 MWh, at least about 1 gigawatt hour (GWh), at least about 5GWh, at least about 10 GWh, at least about 20 GWh, at least 50 GWh, atleast about 100 GWh, at least about 200 GWh, at least about 500 GWh, atleast about 700 GWh, at least about 1000 GWh, or more.

In some cases, a given power capacity may be achieved with a given size,configuration and/or operating conditions of the heat engine/heat pumpcycle. For example, size of turbomachinery, ducts, heat exchangers, orother system components may correspond to a given power capacity.

In some implementations, a given energy storage capacity may be achievedwith a given size and/or number of hot side thermal storage tanks and/orcold side thermal storage tanks. For example, the heat engine/heat pumpcycle can operate at a given power capacity for a given amount of timeset by the heat storage capacity of the system or plant. The numberand/or heat storage capacity of the hot side thermal storage tanks maybe different from the number and/or heat storage capacity of the coldside thermal storage tanks. The number of tanks may depend on the sizeof individual tanks. The size of hot side storage tanks may differ fromthe size of cold side thermal storage tanks. In some cases, the hot sidethermal storage tanks, the hot side heat exchanger and the hot sidethermal storage medium may be referred to as a hot side heat (thermal)storage unit. In some cases, the cold side thermal storage tanks, thecold side heat exchanger and the cold side thermal storage medium may bereferred to as a cold side heat (thermal) storage unit.

A pumped thermal storage facility can include any suitable number of hotside storage tanks, such as at least about 2, at least about 4, at leastabout 10, at least about 50, at least about 100, at least about 500, atleast about 1,000, at least about 5,000, at least about 10,000, and thelike. In some examples, a pumped thermal storage facility includes 2, 3,4, 5, 6, 7, 8, 9, 10, 15, 20, 30, 40, 50, 60, 70, 80, 90, 100, 200, 300,400, 500, 600, 700, 800, 900, 1,000 or more hot side tanks.

A pumped thermal storage facility can also include any suitable numberof cold side storage tanks, such as at least about 2, at least about 4,at least about 10, at least about 50, at least about 100, at least about500, at least about 1,000, at least about 5,000, at least about 10,000,and the like. In some examples, a pumped thermal storage facilityincludes 2, 3, 4, 5, 6, 7, 8, 9, 10, 15, 20, 30, 40, 50, 60, 70, 80, 90,100, 200, 300, 400, 500, 600, 700, 800, 900, 1,000 or more cold sidetanks.

Pumped Thermal Storage Cycles

An aspect of the disclosure relates to pumped thermal systems operatingon pumped thermal storage cycles. In some examples, the cycles allowelectricity to be stored as heat (e.g., in the form of a temperaturedifferential) and then converted back to electricity through the use ofat least two pieces of turbomachinery, a compressor and a turbine. Thecompressor consumes work and raises the temperature and pressure of aworking fluid (WF). The turbine produces work and lowers the temperatureand pressure of the working fluid. In some examples, more than onecompressor and more than one turbine is used. In some cases, the systemcan include at least 1, at least 2, at least 3, at least 4, or at least5 compressors. In some cases, the system can include at least 1, atleast 2, at least 3, at least 4, or at least 5 turbines. The compressorsmay be arranged in series or in parallel. The turbines may be arrangedin series or in parallel.

FIGS. 2A and 2B are schematic flow diagrams of working fluid and heatstorage media of an exemplary pumped thermal system in a charge/heatpump mode and in a discharge/heat engine mode, respectively. The systemmay be idealized for simplicity of explanation so that there are nolosses (i.e., entropy generation) in either the turbomachinery or heatexchangers. The system can include a working fluid 20 (e.g., argon gas)flowing in a closed cycle between a compressor 1, a hot side heatexchanger 2, a turbine 3 and a cold side heat exchanger 4. Fluid flowpaths/directions for the working fluid 20 (e.g., a gas), a hot sidethermal storage (HTS) medium 21 (e.g., a low viscosity liquid) and acold side thermal storage (CTS) medium 22 (e.g., a low viscosity liquid)are indicated by arrows.

FIGS. 3A and 3B are schematic pressure and temperature diagrams of theworking fluid 20 as it undergoes the charge cycles in FIGS. 2A and 2B,respectively, once again simplified in the approximation of no entropygeneration. Normalized pressure is shown on the y-axes and temperatureis shown on the x-axes. The direction of processes taking place duringthe cycles is indicated with arrows, and the individual processes takingplace in the compressor 1, the hot side CFX 2, the turbine 3 and thecold side CFX 4 are indicated on the diagram with their respectivenumerals.

The heat exchangers 2 and 4 can be configured as counter-flow heatexchangers (CFXs), where the working fluid flows in one direction andthe substance it is exchanging heat with is flowing in the oppositedirection. In an ideal counter-flow heat exchanger with correctlymatched flows (i.e., balanced capacities or capacity flow rates), thetemperatures of the working fluid and thermal storage medium flip (i.e.,the counter-flow heat exchanger can have unity effectiveness).

The counter-flow heat exchangers 2 and 4 can be designed and/or operatedto reduce entropy generation in the heat exchangers to negligible levelscompared to entropy generation associated with other system componentsand/or processes (e.g., compressor and/or turbine entropy generation).In some cases, the system may be operated such that entropy generationin the system is minimized. For example, the system may be operated suchthat entropy generation associated with heat storage units is minimized.In some cases, a temperature difference between fluid elementsexchanging heat can be controlled during operation such that entropygeneration in hot side and cold side heat storage units is minimized. Insome instances, the entropy generated in the hot side and cold side heatstorage units is negligible when compared to the entropy generated bythe compressor, the turbine, or both the compressor and the turbine. Insome instances, entropy generation associated with heat transfer in theheat exchangers 2 and 4 and/or entropy generation associated withoperation of the hot side storage unit, the cold side storage unit orboth the hot side and cold side storage units can be less than about50%, less than about 25%, less than about 20%, less than about 15%, lessthan about 10%, less than about 5%, less than about 4%, less than about3%, less than about 2%, or less than about 1% of the total entropygenerated within the system (e.g., entropy generated by the compressor1, the hot side heat exchanger 2, the turbine 3, the cold side heatexchanger 4 and/or other components described herein, such as, forexample, a recuperator). For example, entropy generation can be reducedor minimized if the two substances exchanging heat do so at a localtemperature differential ΔT→0 (i.e., when the temperature differencebetween any two fluid elements that are in close thermal contact in theheat exchanger is small). In some examples, the temperature differentialΔT between any two fluid elements that are in close thermal contact maybe less than about 300 Kelvin (K), less than about 200 K, less thanabout 100 K, less than about 75 K, less than about 50 K, less than about40 K, less than about 30 K, less than about 20 K, less than about 10 K,less than about 5 K, less than about 3 K, less than about 2 K, or lessthan about 1 K. In another example, entropy generation associated withpressure drop can be reduced or minimized by suitable design. In someexamples, the heat exchange process can take place at a constant ornear-constant pressure. Alternatively, a non-negligible pressure dropmay be experienced by the working fluid and/or one or more thermalstorage media during passage through a heat exchanger. Pressure drop inheat exchangers may be controlled (e.g., reduced or minimized) throughsuitable heat exchanger design. In some examples, the pressure dropacross each heat exchanger may be less than about 20% of inlet pressure,less than about 10% of inlet pressure, less than about 5% of inletpressure, less than about 3% of inlet pressure, less than about 2% ofinlet pressure, less than about 1% of inlet pressure, less than about0.5% of inlet pressure, less than about 0.25% of inlet pressure, or lessthan about 0.1% of inlet pressure.

Upon entering the heat exchanger 2, the temperature of the working fluidcan either increase (taking heat from the HTS medium 21, correspondingto the discharge mode in FIGS. 2B and 3B) or decrease (giving heat tothe HTS medium 21, corresponding to the charge mode in FIGS. 2A and 3A),depending on the temperature of the HTS medium in the heat exchangerrelative the temperature of the working fluid. Similarly, upon enteringthe heat exchanger 4, the temperature of the working fluid can eitherincrease (taking heat from the CTS medium 22, corresponding to thecharge mode in FIGS. 2A and 3A) or decrease (giving heat to the CTSmedium 22, corresponding to the discharge mode in FIGS. 2B and 3B),depending on the temperature of the CTS medium in the heat exchangerrelative the temperature of the working fluid.

As described in more detail with reference to the charge mode in FIGS.2A and 3A, the heat addition process in the cold side CFX 4 can takeplace over a different range of temperatures than the heat removalprocess in the hot side CFX 2. Similarly, in the discharge mode in FIGS.2B and 3B, the heat rejection process in the cold side CFX 4 can takeplace over a different range of temperatures than the heat additionprocess in the hot side CFX 2. At least a portion of the temperatureranges of the hot side and cold side heat exchange processes may overlapduring charge, during discharge, or during both charge and discharge.

As used herein, the temperatures T₀, T₁, T₀ ⁺ and T₁ ⁺ are so namedbecause T₀ ⁺, T₁ ⁺ are the temperatures achieved at the exit of acompressor with a given compression ratio r, adiabatic efficiency η_(c)and inlet temperatures of T₀, T₁ respectively. The examples in FIGS. 2A,2B, 3A and 3B can be idealized examples where η_(c)=1 and whereadiabatic efficiency of the turbine η_(t) also has the value η_(t)=1.

With reference to the charge mode shown in FIGS. 2A and 3A, the workingfluid 20 can enter the compressor 1 at position 30 at a pressure P and atemperature T (e.g., at T₁, P₂). As the working fluid passes through thecompressor, work W₁ is consumed by the compressor to increase thepressure and temperature of the working fluid (e.g., to T₁ ⁺, P₁), asindicated by P↑ and T↑ at position 31. In the charge mode, thetemperature T₁ ⁺ of the working fluid exiting the compressor andentering the hot side CFX 2 at position 31 is higher than thetemperature of the HTS medium 21 entering the hot side CFX 2 at position32 from a second hot side thermal storage tank 7 at a temperature T₀ ⁺(i.e., T₀ ⁺<T₁ ⁺). As these two fluids pass in thermal contact with eachother in the heat exchanger, the working fluid's temperature decreasesas it moves from position 31 position 34, giving off heat Q₁ to the HTSmedium, while the temperature of the HTS medium in turn increases as itmoves from position 32 to position 33, absorbing heat Q₁ from theworking fluid. In an example, the working fluid exits the hot side CFX 2at position 34 at the temperature T₀ ⁺ and the HTS medium exits the hotside CFX 2 at position 33 into a first hot side thermal storage tank 6at the temperature T₁ ⁺. The heat exchange process can take place at aconstant or near-constant pressure such that the working fluid exits thehot side CFX 2 at position 34 at a lower temperature but same pressureP₁, as indicated by P and T↓ at position 34. Similarly, the temperatureof the HTS medium 21 increases in the hot side CFX 2, while its pressurecan remain constant or near-constant.

Upon exiting the hot side CFX 2 at position 34 (e.g., at T₀ ⁺, P₁), theworking fluid 20 undergoes expansion in the turbine 3 before exiting theturbine at position 35. During the expansion, the pressure andtemperature of the working fluid decrease (e.g., to T₀, P₂), asindicated by P↓ and T↓ at position 35. The magnitude of work W₂generated by the turbine depends on the enthalpy of the working fluidentering the turbine and the degree of expansion. In the charge mode,heat is removed from the working fluid between positions 31 and 34 (inthe hot side CFX 2) and the working fluid is expanded back to thepressure at which it initially entered the compressor at position 30(e.g., P₂). The compression ratio (e.g., P₁/P₂) in the compressor 1being equal to the expansion ratio in the turbine 3, and the enthalpy ofthe gas entering the turbine being lower than the enthalpy of the gasexiting the compressor, the work W₂ generated by the turbine 3 issmaller than the work W₁ consumed by the compressor 1 (i.e., W₂<W₁).

Because heat was taken out of the working fluid in the hot side CFX 2,the temperature T₀ at which the working fluid exits the turbine atposition 35 is lower than the temperature T₁ at which the working fluidinitially entered the compressor at position 30. To close the cycle(i.e., to return the pressure and temperature of the working fluid totheir initial values T₁, P₂ at position 30), heat Q₂ is added to theworking fluid from the CTS medium 22 in the cold side CFX 4 betweenpositions 35 and 30 (i.e., between the turbine 3 and the compressor 1).In an example, the CTS medium 22 enters the cold side CFX 4 at position36 from a first cold side thermal storage tank 8 at the temperature T₁and exits the cold side CFX 4 at position 37 into a second cold sidethermal storage tank 9 at the temperature T₀, while the working fluid 20enters the cold side CFX 4 at position 35 at the temperature T₀ andexits the cold side CFX 4 at position 30 at the temperature T₁. Again,the heat exchange process can take place at a constant or near-constantpressure such that the working fluid exits the cold side CFX 2 atposition 30 at a higher temperature but same pressure P₂, as indicatedby P and T↑ at position 30. Similarly, the temperature of the CTS medium22 decreases in the cold side CFX 2, while its pressure can remainconstant or near-constant.

During charge, the heat Q₂ is removed from the CTS medium and the heatQ₁ is added to the HTS medium, wherein Q₁>Q₂. A net amount of work W₁−W₂is consumed, since the work W₁ used by the compressor is greater thanthe work W₂ generated by the turbine. A device that consumes work whilemoving heat from a cold body or thermal storage medium to a hot body orthermal storage medium is a heat pump; thus, the pumped thermal systemin the charge mode operates as a heat pump.

In an example, the discharge mode shown in FIGS. 2B and 3B can differfrom the charge mode shown in FIGS. 2A and 3A in the temperatures of thethermal storage media being introduced into the heat exchangers. Thetemperature at which the HTS medium enters the hot side CFX 2 atposition 32 is T₁ ⁺ instead of T₀ ⁺, and the temperature of the CTSmedium entering the cold side CFX 4 at position 36 is T₀ instead of T₁.During discharge, the working fluid enters the compressor at position 30at T₀ and P₂, exits the compressor at position 31 at T₀ ⁺<T₁ ⁺ and P₁,absorbs heat from the HTS medium in the hot side CFX 2, enters theturbine 3 at position 34 at T₁ ⁺ and P₁, exits the turbine at position35 at T₁>T₀ and P₂, and finally rejects heat to the CTS medium in thecold side CFX 4, returning to its initial state at position 30 at T₀ andP₂.

The HTS medium at temperature T₁ ⁺ can be stored in a first hot sidethermal storage tank 6, the HTS medium at temperature T₀ ⁺ can be storedin a second hot side thermal storage tank 7, the CTS medium attemperature T₁ can be stored in a first cold side thermal storage tank8, and the CTS medium at temperature T₀ can be stored in a second coldside thermal storage tank 9 during both charge and discharge modes. Inone implementation, the inlet temperature of the HTS medium at position32 can be switched between T₁ ⁺ and T₀ ⁺ by switching between tanks 6and 7, respectively. Similarly, the inlet temperature of the CTS mediumat position 36 can be switched between T₁ and T₀ by switching betweentanks 8 and 9, respectively. Switching between tanks can be achieved byincluding a valve or a system of valves (e.g., valve systems 12 and 13in FIG. 4B) for switching connections between the hot side heatexchanger 2 and the hot side tanks 6 and 7, and/or between the cold sideheat exchanger 4 and the cold side tanks 8 and 9 as needed for thecharge and discharge modes. In some implementations, connections may beswitched on the working fluid side instead, while the connections ofstorage tanks 6, 7, 8 and 9 to the heat exchangers 2 and 4 remainstatic. In some examples, flow paths and connections to the heatexchangers may depend on the design (e.g., shell-and-tube) of each heatexchanger. In some implementations, one or more valves can be used toswitch the direction of both the working fluid and the heat storagemedium through the counter-flow heat exchanger on charge and discharge.Such configurations may be used, for example, due to high thermalstorage capacities of the heat exchanger component, to decrease oreliminate temperature transients, or a combination thereof. In someimplementations, one or more valves can be used to switch the directionof only the working fluid, while the direction of the HTS or CTS can bechanged by changing the direction of pumping, thereby maintaining thecounter-flow configuration. In some implementations, different valveconfigurations may be used for the HTS and the CTS. Further, anycombination of the valve configurations herein may be used. For example,the system may be configured to operate using different valveconfigurations in different situations (e.g., depending on systemoperating conditions).

In the discharge mode shown in FIGS. 2B and 3B, the working fluid 20 canenter the compressor 1 at position 30 at a pressure P and a temperatureT (e.g., at T₀, P₂). As the working fluid passes through the compressor,work W₁ is consumed by the compressor to increase the pressure andtemperature of the working fluid (e.g., to T₀ ⁺, P₁), as indicated by P↑and T↑ at position 31. In the discharge mode, the temperature T₀ ⁺ ofthe working fluid exiting the compressor and entering the hot side CFX 2at position 31 is lower than the temperature of the HTS medium 21entering the hot side CFX 2 at position 32 from a first hot side thermalstorage tank 6 at a temperature T₁ ⁺ (i.e., T₀ ⁺<T₁ ⁺). As these twofluids pass in thermal contact with each other in the heat exchanger,the working fluid's temperature increases as it moves from position 31position 34, absorbing heat Q₁ from the HTS medium, while thetemperature of the HTS medium in turn decreases as it moves fromposition 32 to position 33, giving off heat Q₁ to the working fluid. Inan example, the working fluid exits the hot side CFX 2 at position 34 atthe temperature T₁ ⁺ and the HTS medium exits the hot side CFX 2 atposition 33 into the second hot side thermal storage tank 7 at thetemperature T₀ ⁺. The heat exchange process can take place at a constantor near-constant pressure such that the working fluid exits the hot sideCFX 2 at position 34 at a higher temperature but same pressure P₁, asindicated by P and T↑ at position 34. Similarly, the temperature of theHTS medium 21 decreases in the hot side CFX 2, while its pressure canremain constant or near-constant.

Upon exiting the hot side CFX 2 at position 34 (e.g., at T₁ ⁺, P₁), theworking fluid 20 undergoes expansion in the turbine 3 before exiting theturbine at position 35. During the expansion, the pressure andtemperature of the working fluid decrease (e.g., to T₁, P₂), asindicated by P↓ and T↓ at position 35. The magnitude of work W₂generated by the turbine depends on the enthalpy of the working fluidentering the turbine and the degree of expansion. In the discharge mode,heat is added to the working fluid between positions 31 and 34 (in thehot side CFX 2) and the working fluid is expanded back to the pressureat which it initially entered the compressor at position 30 (e.g., P₂).The compression ratio (e.g., P₁/P₂) in the compressor 1 being equal tothe expansion ratio in the turbine 3, and the enthalpy of the gasentering the turbine being higher than the enthalpy of the gas exitingthe compressor, the work W₂ generated by the turbine 3 is greater thanthe work W₁ consumed by the compressor 1 (i.e., W₂>W₁).

Because heat was added to the working fluid in the hot side CFX 2, thetemperature T₁ at which the working fluid exits the turbine at position35 is higher than the temperature T₀ at which the working fluidinitially entered the compressor at position 30. To close the cycle(i.e., to return the pressure and temperature of the working fluid totheir initial values T₀, P₂ at position 30), heat Q₂ is rejected by theworking fluid to the CTS medium 22 in the cold side CFX 4 betweenpositions 35 and 30 (i.e., between the turbine 3 and the compressor 1).The CTS medium 22 enters the cold side CFX 4 at position 36 from asecond cold side thermal storage tank 9 at the temperature T₀ and exitsthe cold side CFX 4 at position 37 into a first cold side thermalstorage tank 8 at the temperature T₁, while the working fluid 20 entersthe cold side CFX 4 at position 35 at the temperature T₁ and exits thecold side CFX 4 at position 30 at the temperature T₀. Again, the heatexchange process can take place at a constant or near-constant pressuresuch that the working fluid exits the cold side CFX 2 at position 30 ata higher temperature but same pressure P₂, as indicated by P and T↓ atposition 30. Similarly, the temperature of the CTS medium 22 increasesin the cold side CFX 2, while its pressure can remain constant ornear-constant.

During discharge, the heat Q₂ is added to the CTS medium and the heat Q₁is removed from the HTS medium, wherein Q₁>Q₂. A net amount of workW₂−W₁ is generated, since the work W₁ used by the compressor is smallerthan the work W₂ generated by the turbine. A device that generates workwhile moving heat from a hot body or thermal storage medium to a coldbody or thermal storage medium is a heat engine; thus, the pumpedthermal system in the discharge mode operates as a heat engine.

FIG. 4A is a simplified schematic perspective view of a closed workingfluid system in the pumped thermal system in FIGS. 2A-2B. As indicated,the working fluid 20 (contained inside tubing) circulates clockwisebetween the compressor 1, the hot side heat exchanger 2, the turbine 3,and the cold side heat exchanger 4. The compressor 1 and the turbine 3can be ganged on a common mechanical shaft 10 such that they rotatetogether. In some implementations, the compressor 1 and the turbine 3can have separate mechanical shafts. A motor/generator 11 (e.g.,including a synchronous motor-synchronous generator converter on asingle common shaft) provides power to and from the turbomachinery. Inthis example, the compressor, the turbine and the motor/generator areall located on a common shaft. Pipes at positions 32 and 33 transfer hotside thermal storage fluid to and from the hot side heat exchanger 2,respectively. Pipes at positions 36 and 37 transfer cold side thermalstorage fluid to and from the cold side heat exchanger 4, respectively.

Although the system of FIG. 4A is illustrated as comprising a compressor1 and turbine 3, the system can include one or more compressors and oneor more turbines, which may operate, for example, in a parallelconfiguration, or alternatively in a series configuration or in acombination of parallel and series configurations. In some examples, asystem of compressors or turbines may be assembled such that a givencompression ratio is achieved. In some cases, different compressionratios (e.g., on charge and discharge) can be used (e.g., by connectingor disconnecting, in a parallel and/or series configuration, one or morecompressors or turbines from the system of compressors or turbines). Insome examples, the working fluid is directed to a plurality ofcompressors and/or a plurality of turbines. In some examples, thecompressor and/or turbine may have temperature dependent compressionratios. Arrangement and/or operation of the turbomachinery and/or otherelements of the system may be adjusted in accordance with thetemperature dependence (e.g., to optimize performance).

FIG. 4B is a simplified schematic perspective view of the pumped thermalsystem in FIGS. 2A-2B with hot side and cold side storage tanks and aclosed cycle working fluid system. In this example, the HTS medium is amolten salt and the CTS medium is a low temperature liquid. One, two ormore first hot side tanks 6 (at the temperature T₁ ⁺) and one, two ormore second hot side tanks 7 (at the temperature T₀ ⁺), both for holdingthe HTS medium, are in fluid communication with a valve 13 configured totransfer the HTS medium to and from the hot side heat exchanger 2. One,two or more first cold side tanks 8 (at the temperature T₁) and one, twoor more second cold side tanks 9 (at the temperature T₀), both forholding the CTS medium, are in fluid communication with a valve 12configured to transfer the CTS medium to and from the cold side heatexchanger 4.

The thermal energy reservoirs or storage tanks may be thermallyinsulated tanks that can hold a suitable quantity of the relevantthermal storage medium (e.g., heat storage fluid). The storage tanks mayallow for relatively compact storage of large amounts of thermal energy.In an example, the hot side tanks 6 and/or 7 can have a diameter ofabout 80 meters, while the cold side tanks 8 and/or 9 can have adiameter of about 60 meters. In another example, the size of each (i.e.,hot side or cold side) thermal storage for a 1 GW plant operating for 12hours can be about 20 medium-sized oil refinery tanks.

In some implementations, a third set of tanks containing storage mediaat intermediate temperatures between the other tanks may be included onthe hot side and/or the cold side. In an example, a third storage ortransfer tank (or set of tanks) at a temperature intermediate to thetemperatures of a first tank (or set of tanks) and a second tank (or setof tanks) may be provided. A set of valves may be provided for switchingthe storage media between the different tanks and heat exchangers. Forexample, thermal media may be directed to different sets of tanks afterexiting the heat exchangers depending on operating conditions and/orcycle being used. In some implementations, one or more additional setsof storage tanks at different temperatures may be added on the hot sideand/or the cold side.

The storage tanks (e.g., hot side tanks comprising hot side thermalstorage medium and/or cold side tanks comprising cold side thermalstorage medium) may operate at ambient pressure. In someimplementations, thermal energy storage at ambient pressure can providesafety benefits. Alternatively, the storage tanks may operate atelevated pressures, such as, for example, at a pressure of at leastabout 2 atm, at least about 5 atm, at least about 10 atm, at least about20 atm, or more. Alternatively, the storage tanks may operate at reducedpressures, such as, for example, at a pressure of at most about 0.9 atm,at most about 0.7 atm, at most about 0.5 atm, at most about 0.3 atm, atmost about 0.1 atm, at most about 0.01 atm, at most about 0.001 atm, orless. In some cases (e.g., when operating at higher/elevated or lowerpressures or to avoid contamination of the thermal storage media), thestorage tanks can be sealed from the surrounding atmosphere.Alternatively, in some cases, the storage tanks may not be sealed. Insome implementations, the tanks may include one or more pressureregulation or relief systems (e.g., a valve for safety or systemoptimization).

As used herein, the first hot side tank(s) 6 (at the temperature T₁ ⁺)can contain HTS medium at a higher temperature than the second hot sidetank(s) 7 (at the temperature T₀ ⁺), and the first cold side tank(s) 8(at the temperature T₁) can contain CTS medium at a higher temperaturethan the second cold side tank(s) 9 (at the temperature T₀). Duringcharge, HTS medium in the first (higher temperature) hot side tank(s) 6and/or CTS medium in the second (lower temperature) cold side tank(s) 9can be replenished. During discharge, HTS medium in the first (highertemperature) hot side tank(s) 6 and/or CTS medium in the second (lowertemperature) cold side tank(s) 9 can be consumed.

In the foregoing examples, in either mode of operation, two of the fourstorage tanks 6, 7, 8 and 9 are feeding thermal storage medium to theheat exchangers 2 and 4 at the inlets 32 and 36, respectively, and theother two tanks are receiving thermal storage medium from the heatexchangers 2 and 4 from the exits 33 and 37, respectively. In thisconfiguration, the feed tanks can contain a storage medium at a giventemperature due to prior operating conditions, while the receivingtanks' temperatures can depend on current system operation (e.g.,operating parameters, loads and/or power input). The receiving tanktemperatures may be set by the Brayton cycle conditions. In some cases,the receiving tank temperatures may deviate from desired values due todeviations from predetermined cycle conditions (e.g., variation ofabsolute pressure in response to system demand) and/or due to entropygeneration within the system. In some cases (e.g., due to entropygeneration), at least one of the four tank temperatures can be higherthan desired. In some implementations, a radiator can be used to rejector dissipate this waste heat to the environment. In some cases, heatrejection to the environment may be enhanced (e.g., using evaporativecooling etc.). The waste heat generated during operation of the pumpedthermal systems herein can also be utilized for other purposes. Forexample, waste heat from one part of the system may be used elsewhere inthe system. In another example, waste heat may be provided to anexternal process or system, such as, for example, a manufacturingprocess requiring low grade heat, commercial or residential heating,thermal desalination, commercial drying operations etc.

Components of pumped thermal systems of the disclosure may exhibitnon-ideal performance, leading to losses and/or inefficiencies. Themajor losses in the system may occur due to inefficiencies of theturbomachinery (e.g., compressor and turbine) and the heat exchangers.The losses due to the heat exchangers may be small compared to thelosses due to the turbomachinery. In some implementations, the lossesdue to the heat exchangers can be reduced to near zero with suitabledesign and expense. Therefore, in some analytical examples, losses dueto the heat exchangers and other possible small losses due to pumps, themotor/generator and/or other factors may be neglected.

Losses due to turbomachinery can be quantified in terms of adiabaticefficiencies η_(c) and η_(t) (also known as isentropic efficiencies) forcompressors and turbines, respectively. For large turbomachinery,typical values may range between η_(c)=0.85-0.9 for compressors andη_(t)=0.9-0.95 for turbines. The actual amount of work produced orconsumed by a cycle can then be expressed as

${{\Delta W} = {{W_{actual}^{({out})} - W_{actual}^{({in})}} = {{\eta_{t}W_{ideal}^{({out})}} - {\frac{1}{\eta_{c}}W_{ideal}^{({in})}}}}},$

where, in an example assuming constant specific heats of the workingfluid, W_(ideal) ^((in))=c_(p)T_(inlet)(ψ−1), W_(ideal)^((out))=c_(p)T_(inlet)(1−ψ⁻¹), where

${\psi = r^{\frac{\gamma - 1}{\gamma}}},$

r is the compression ratio (i.e., ratio of the higher pressure to thelower pressure), and γ=c_(p)/c_(v) is the ratio of specific heats of theworking fluid. Due to compressor and turbine inefficiencies, more workis required to achieve a given compression ratio during compression, andless work is generated during expansion for a given compression ratio.Losses can also be quantified in terms of the polytropic, or singlestage, efficiencies, η_(cp) and η_(tp), for compressors and turbines,respectively. The polytropic efficiencies are related to the adiabaticefficiencies η_(c) and η_(t) the equations

$\eta_{c} = {{\frac{\psi - 1}{\psi^{1/\eta_{cp}} - 1}\mspace{14mu} {and}\mspace{14mu} \eta_{t}} = {\frac{1 - \psi^{- \eta_{tp}}}{1 - \psi^{- 1}}.}}$

In examples where η_(c)=η_(t)=1, pumped thermal cycles of the disclosurecan follow identical paths in both charge and discharge cycles (e.g., asshown in FIGS. 3A and 3B). In examples where η_(c)<1 and/or η_(t)<1,compression in the compressor can lead to a greater temperature increasethan in the ideal case for the same compression ratio, and expansion inthe turbine can lead to a smaller temperature decrease than in the idealcase.

In some implementations, the polytropic efficiency of the compressorη_(cp) may be at least about 0.3, at least about 0.5, at least about0.6, at least about 0.7, at least about 0.75, at least about 0.8, atleast about 0.85, at least about 0.9, at least about 0.91, at leastabout 0.92, at least about 0.93, at least about 0.96, or more. In someimplementations, the polytropic efficiency of the compressor η_(tp) maybe at least about 0.3, at least about 0.5, at least about 0.6, at leastabout 0.7, at least about 0.75, at least about 0.8, at least about 0.85,at least about 0.9, at least about 0.91, at least about 0.92, at leastabout 0.93, at least about 0.96, at least about 0.97 or more.

T₀ ⁺, T₁ ⁺ were previously defined as the temperatures achieved at theexit of a compressor with a given compression ratio r, adiabaticefficiency η_(c) and inlet temperatures of T₀, T₁ respectively. In someexamples, these four temperatures are related by the equation

$\frac{T_{0}^{+}}{T_{0}} = {\frac{T_{1}^{+}}{T_{1}} = {\psi^{1/\eta_{cp}}.}}$

FIG. 5A shows an exemplary heat storage charge cycle for a water(CTS)/molten salt (HTS) system with η_(c)=0.9 and η_(t)=0.95. The dashedlines correspond to η_(c)=η_(t)=1 and the solid lines show the chargecycle with η_(t)=0.95 and η_(c)=0.9. In this example, the CTS medium onthe cold side is water, and the HTS medium on the hot side is moltensalt. In some cases, the system can include 4 heat storage tanks. In thecharge cycle, the working fluid at T₀ and P₂ can exchange heat with aCTS medium in the cold side heat exchanger 4, whereby its temperaturecan increase to T₁ (assuming negligible pressure drop, its pressure canremain P₂). In the compressor 1 with η_(c)=0.9, the temperature andpressure of the working fluid can increase from T₁, P₂ to T₁ ⁺, P₁. Theworking fluid can then exchange heat with an HTS medium in the hot sideheat exchanger 2, such that its temperature can decrease (at constantpressure P₁, assuming negligible pressure drop). If the working fluidenters the turbine 3 with η_(t)=0.95 at the temperature T₀ ⁺ and expandsback to its original pressure P₂, its temperature when exiting theturbine may not be T₀. Instead, the working fluid may enter the turbineat a temperature {tilde over (T)}₀ ⁺ and exit the turbine at thetemperature T₀ and pressure P₂. In some examples, the temperatures arerelated by the relation

$\frac{{\overset{\sim}{T}}_{0}^{+}}{T_{0}} = {\psi^{\eta_{tp}}.}$

In some examples, {tilde over (T)}₀ ⁺ is the temperature at which theworking fluid enters the inlet of a turbine with adiabatic efficiencyη_(t) and compression ratio r in order to exit at the temperature T₀.

In some implementations, the temperature {tilde over (T)}₀ ⁺ may beincorporated into charge cycles of the disclosure by first heatexchanging the working fluid with the HTS medium from T₁ ⁺ to T₀ ⁺,followed by further cooling the working fluid from T₀ ⁺ to {tilde over(T)}₀ ⁺ as illustrated by section 38 of the cycle in FIG. 5A.

FIG. 5B shows an exemplary heat storage discharge (extraction) cycle forthe water/molten salt system in FIG. 5A with η_(c)=0.9 and η_(t)=0.95.The dashed lines correspond to η_(c)=η_(t)=1 and the solid lines showthe charge cycle with η_(t)=0.95 and η_(c)=0.9. In the discharge cycle,the working fluid at T₁ and P₂ can exchange heat with a CTS medium inthe cold side heat exchanger 4, whereby its temperature can decrease toT₀ (assuming negligible pressure drop, its pressure can remain P₂). Inthe compressor 1 with η_(c)=0.9, the temperature and pressure of theworking fluid can increase from T₀, P₂ to T₀ ⁺, P₁. The working fluidcan then exchange heat with an HTS medium in the hot side heat exchanger2, such that its temperature can increase (at constant pressure P₁,assuming negligible pressure drop). Working fluid entering the turbine 3at T₁ ⁺ may not exit the turbine at the temperature T₁ as in the chargecycle, but may instead exit at a temperature {tilde over (T)}₁, where,in some examples, {tilde over (T)}₁=T₁ ⁺ψ^(−η) ^(tp) . In some examples,{tilde over (T)}₁ is the temperature at which the working fluid exitsthe outlet of a turbine with adiabatic efficiency η_(t) and compressionratio r after entering the inlet of the turbine at the temperature T₁ ⁺.

In some implementations, the temperature {tilde over (T)}₁ may beincorporated into the discharge cycles of the disclosure by firstcooling the working fluid exiting the turbine at {tilde over (T)}₁ toT₁, as illustrated by section 39 of the cycle in FIG. 5B, followed byheat exchanging the working fluid with the CTS medium form T₁ to T₀.

The charge and discharge cycles may be closed by additional heatrejection operations in sections 38 (between T₀ ⁺ and {tilde over (T)}₀⁺) and 39 (between {tilde over (T)}₁ and T₁), respectively. In somecases, closing the cycles through heat rejection in sections of thecycles where the working fluid can reject heat to ambient at low costmay eliminate the need for additional heat input into the system. Thesections of the cycles where the working fluid can reject heat toambient may be limited to sections where the temperature of the workingfluid is high enough above ambient temperature for ambient cooling to befeasible. In some examples, heat may be rejected to the environment insections 38 and/or 39. For example, heat may be rejected using one ormore working fluid to air radiators, intermediate water cooling, or anumber of other methods. In some cases, heat rejected in sections 38and/or 39 may be used for another useful purpose, such as, for example,cogeneration, thermal desalination and/or other examples describedherein.

In some implementations, the cycles may be closed by varying thecompression ratios between the charge and discharge cycles, as shown,for example, in FIG. 5C. The ability to vary compression ratio on chargeand discharge may be implemented, for example, by varying the rotationspeed of the compressor and/or turbine, by variable stator pressurecontrol, by bypassing a subset of the compression or expansion stages oncharge or discharge by the use of valves, or by using dedicatedcompressor/turbine pairs for charge and discharge mode. In one example,the compression ratio in the discharge cycle in FIG. 5B can be changedsuch that heat rejection in section is 39 is not used, and only heatrejection in section 38 in the charge cycle is used. Varying thecompression ratio may allow heat (i.e., entropy) to be rejected at alower temperature, thereby increasing overall roundtrip efficiency. Insome examples of this configuration, the compression ratio on charge,r_(C), can be set such that

${\frac{T_{1}^{+}}{T_{1}} = \psi_{C}^{1/\eta_{cp}}},$

and on discharge, the compression ratio r_(D) can be set such that

$\frac{T_{1}^{+}}{T_{1}} = {\psi_{D}^{\eta_{tp}}.}$

In some cases, the upper temperatures T₁ ⁺ and T₁ can be identical oncharge and discharge and no heat removal may be needed in this portion(also “leg” herein) of the cycle. In such cases, the temperature T₀ ⁺ oncharge (e.g., T₀ ^(+(c))=T₀ψ_(C) ^(η) ^(cp) and the temperature T₀ ⁺ ondischarge (e.g., T₀ ^(+(d))=T₀ψ_(C) ^(1/η) ^(cp) ) can be dissimilar andheat may be rejected (also “dissipated” or “dumped” herein) to theenvironment between the temperatures T₀ ^(+(c)) and T₀ ^(+(d)). In animplementation where only the storage media exchange heat with theenvironment, a heat rejection device (e.g., devices 55 and 56 shown inFIG. 7D) can be used to lower the temperature of the CTS from T₀ ^(+(d))to T₀ ^(+(c)) between discharge and charge.

FIG. 5C shows an example of a cycle with variable compression ratios.The compression ratio can be higher on discharge (when work is producedby the system) than on charge (when work is consumed by the system),which may increase an overall round trip efficiency of the system. Forexample, during a charge cycle 80 with T₀ ^(+(c)), a lower compressionratio of <3 can be used; during a discharge cycle 81 with T₀ ^(+(d)), acompression ratio of >3 can be used. The upper temperatures reached inboth cycles 80 and 81 can be T₁ and T₁ ⁺, and no excess heat may berejected.

The compression ratio may be varied between charge and discharge suchthat the heat dissipation to the environment needed for closing thecycle on both charge and discharge occurs between the temperatures T₀^(+(C)) (the temperature of the working fluid before it enters theturbine during the charge cycle) and T₀ ^(+(D)) (the temperature of theworking fluid as it exits the compressor on discharge) and not above thetemperature T₁ (the temperature of the working fluid before it entersthe compressor on charge and/or exits the turbine on discharge). In someexamples, none of the heat is rejected at a temperature above the lowesttemperature of the HTS medium.

In the absence of system losses and/or inefficiencies, such as, forexample, in the case of pumped thermal systems comprising heat pump(s)and heat engine(s) operating at the Carnot limit, a given amount of heatQ_(H) can be transferred using a given quantity of work W in heat pump(charge) mode, and the same Q_(H) can be used in heat engine (discharge)mode to produce the same work W, leading to a unity (i.e., 100%)roundtrip efficiency. In the presence of system losses and/orinefficiencies, roundtrip efficiencies of pumped thermal systems may belimited by how much the components deviate from ideal performance.

The roundtrip efficiency of a pumped thermal system may be defined asη_(store)=|W_(cv) ^(extract)|/|W_(cv) ^(charge)|. In some examples, withan approximation of ideal heat exchange, the roundtrip efficiency can bederived by considering the net work output during the discharge cycle,

${{W_{cv}^{extract}} = {{\eta_{t}W_{ideal}^{out}} - \frac{W_{ideal}^{in}}{\eta_{c}}}},$

and the net work input during the charge cycle,

${W_{cv}^{charge}} = {\frac{W_{ideal}^{in}}{\eta_{c}} - {\eta_{t}W_{ideal}^{out}}}$

using the equations for work and temperature given above.

Roundtrip efficiencies may be calculated for different configurations ofpumped thermal systems (e.g., for different classes of thermal storagemedia) based on turbomachinery component efficiencies, η_(c) and η_(t).

In one example, FIG. 6A shows roundtrip efficiency contours for awater/salt system, such as, for example, the water/salt system in FIGS.5A and 5B with T₀=273 K (0° C.), T₁=373 K (100° C.) and a compressionratio of r=5.65 chosen to achieve compatibility with the salt(s) on thehot side. Exemplary roundtrip efficiency contours at values of η_(store)of 10%, 20%, 30%, 40%, 50%, 60%, 70%, 80% and 90% are shown as afunction of component efficiencies η_(c) and η_(t) on the x- and y-axes,respectively. The symbols ⊕ and

represent the approximate range of present large turbomachineryadiabatic efficiency values. The dashed arrows represent the directionof increasing efficiency.

FIG. 6B shows roundtrip efficiency contours for a colder storage/saltsystem, such as, for example a hexane/salt system with a gas-gas heatexchanger in FIGS. 7A, 7B, 8A and 8B with T₀=194 K (−79° C.), T₁=494 K(221° C.) and a compression ratio of r=3.28. Exemplary roundtripefficiency contours at values of η_(store) of 10%, 20%, 30%, 40%, 50%,60%, 70%, 80% and 90% are shown as a function of component efficienciesη_(c) and η_(t) on the x- and y-axes, respectively. The symbols ⊕ and

represent the approximate range of present large turbomachineryadiabatic efficiency values. As discussed in detail elsewhere herein,using hexane, heptane and/or another CTS medium capable of lowtemperature operation may result in significant improvements of systemefficiency.

Pumped Thermal Storage Cycles with Recuperation

Another aspect of the disclosure is directed to pumped thermal systemswith regeneration/recuperation. In some situations, the termsregeneration and recuperation can be used interchangeably, although theymay have different meanings. As used herein, the terms “recuperation”and “recuperator” generally refer to the presence of one or moreadditional heat exchangers where the working fluid exchanges heat withitself during different segments of a thermodynamic cycle throughcontinuous heat exchange without intermediate thermal storage. As usedherein, the terms “regeneration” and “regenerator” may be used todescribe the same configuration as the terms “recuperation” and“recuperator.” The roundtrip efficiency of pumped thermal systems may besubstantially improved if the allowable temperature ranges of thestorage materials can be extended. In some implementations, this may beaccomplished by choosing a material or medium on the cold side that cango to temperatures below 273 K (0° C.). For example, a CTS medium (e.g.,hexane) with a low temperature limit of approximately T₀=179 K (−94° C.)may be used in a system with a molten salt HTS medium. However, T₀ ⁺(i.e., the lowest temperature of the working fluid in the hot side heatexchanger) at some (e.g., modest) compression ratios may be below thefreezing point of the molten salt, making the molten salt unviable asthe HTS medium. In some implementations, this can be resolved byincluding a working fluid to working fluid (e.g., gas-gas) heatexchanger (also “recuperator” or “regenerator” herein) in the cycle.

FIG. 7A is a schematic flow diagram of working fluid and heat storagemedia of a pumped thermal system in a charge/heat pump mode with agas-gas heat exchanger 5 for the working fluid. The use of the gas-gasheat exchanger can enable use of colder heat storage medium on the coldside of the system. The working fluid can be argon. The working fluidcan be a mixture of primarily argon mixed with another gas such ashelium. For example, the working fluid may comprise at least about 50%argon, at least about 60% argon, at least about 70% argon, at leastabout 80% argon, at least about 90% argon, or about 100% argon, withbalance helium.

FIG. 8A shows a heat storage charge cycle for the storage system in FIG.7A with a cold side storage medium (e.g., liquid hexane) capable ofgoing down to approximately to 179 K (−94° C.) and a molten salt as thehot side storage, and η_(c)=0.9 and η_(t)=0.95. The CTS medium can behexane or heptane and the HTS medium can be molten salt. In some cases,the system can include four heat storage tanks.

In one implementation, during charge in FIGS. 7A and 8A, the workingfluid enters the compressor at T₁ and P₂, exits the compressor at T₁ ⁺and P₁, rejects heat Q₁ to the HTS medium 21 in the hot side CFX 2,exiting the hot side CFX 2 at T₁ and P₁, rejects heat Q_(recup) (also“Q_(regen)” herein, as shown, for example, in the accompanying drawings)to the cold (low pressure) side working fluid in the heat exchanger orrecuperator 5, exits the recuperator 5 at T₀ ⁺ and P₁, rejects heat tothe environment (or other heat sink) in section 38 (e.g., a radiator),enters the turbine 3 at {tilde over (T)}₀ ⁺ and P₁, exits the turbine atT₀ and P₂, absorbs heat Q₂ from the CTS medium 22 in the cold side CFX4, exiting the cold side CFX 4 at T₀ ⁺ and P₂, absorbs heat Q_(recup)from the hot (high pressure) side working fluid in the heat exchanger orrecuperator 5, and finally exits the recuperator 5 at T₁ and P₂,returning to its initial state before entering the compressor.

FIG. 7B is a schematic flow diagram of working fluid and heat storagemedia of the pumped thermal system in FIG. 7A in a discharge/heat enginemode. Again, the use of the gas-gas heat exchanger can enable use ofcolder heat storage fluid (CTS) and/or colder working fluid on the coldside of the system.

FIG. 8B shows a heat storage discharge cycle for the storage system forthe storage system in FIG. 7B with a cold side storage medium (e.g.,liquid hexane) capable of going down to 179 K (−94° C.) and a moltensalt as the hot side storage, and η_(c)=0.9 and η_(t)=0.95. Again, theCTS medium can be hexane or heptane and the HTS medium can be moltensalt, and the system may include 4 heat storage tanks.

During discharge in FIGS. 7B and 8B, the working fluid enters thecompressor at T₀ and P₂, exits the compressor at T₀ ⁺ and P₁, absorbsheat Q_(recup) from the cold (low pressure) side working fluid in theheat exchanger or recuperator 5, exits the recuperator 5 at T₁ and P₁,absorbs heat Q₁ from the HTS medium 21 in the hot side CFX 2, exitingthe hot side CFX 2 at T₁ ⁺ and P₁, enters the turbine 3 at T₁ ⁺ and P₁,exits the turbine at {tilde over (T)}₁ and P₂, rejects heat to theenvironment (or other heat sink) in section 39 (e.g., a radiator),rejects heat Q_(recup) to the hot (high pressure) side working fluid inthe heat exchanger or recuperator 5, enters the cold side CFX 4 at T₀ ⁺and P₂, rejects heat Q₂ to the CTS medium 22 in the cold side CFX 4, andfinally exits the cold side CFX 4 at T₀ and P₂, returning to its initialstate before entering the compressor.

In another implementation, shown in FIG. 7C, the charge cycle remainsthe same as in FIGS. 7A and 8A, except that the working fluid exits therecuperator 5 at {tilde over (T)}₀ ⁺ and P₁ (instead of at T₀ ⁺ and P₁as in FIGS. 7A and 8A), enters the turbine 3 at {tilde over (T)}₀ ⁺ andP₁, exits the turbine at T₀ and P₂, absorbs heat Q₂ from the CTS medium22 having a temperature {tilde over (T)}₀ ⁺ (instead of at T₀ ⁺ as inFIGS. 7A and 8A) in the cold side CFX 4, and exits the cold side CFX 4at {tilde over (T)}₀ ⁺ and P₂ (instead of at T₀ ⁺ and P₂ as in FIG. 7A)before reentering the recuperator 5. The heat between temperatures T₀ ⁺and {tilde over (T)}₀ ⁺ is no longer rejected from the working fluid tothe environment directly (as in section 38 in FIGS. 7A and 8A).

During discharge in FIG. 7D, the discharge cycle remains the same as inFIGS. 7B and 8B, except that the temperature of the HTS medium beingdeposited in tank 7 is changed. The working fluid exits the recuperator5 at {tilde over (T)}₁ and P₁ (instead of at T₁ and P₁ as in FIGS. 7Band 8B) and absorbs heat Q₁ from the HTS medium 21 in the hot side CFX2. The HTS medium exits the hot side CFX 2 having a temperature {tildeover (T)}₁ (instead of at T₁ as in FIGS. 7B and 8B). The working fluidthen exits the hot side CFX 2 at T₁ ⁺ and P₁, enters the turbine 3 at T₁⁺ and P₁, and exits the turbine at {tilde over (T)}₁ and P₂ beforereentering the recuperator 5. The heat between temperatures {tilde over(T)}₁ and T₁ is no longer rejected from the working fluid to theenvironment directly (as in section 39 in FIGS. 7B and 8B). As in FIG.7B, the CTS medium enters the tank 8 at temperature T₀ ⁺.

After the discharge in FIG. 7D, in preparation for the charge in FIG.7C, heat exchange with ambient may be used to cool the HTS medium 21from the temperature {tilde over (T)}₁ used in the discharge cycle tothe temperature T₁ used in the charge cycle. Similarly, heat exchangewith ambient may be used to cool the CTS medium 22 from the temperatureT₀ ⁺ used in the discharge cycle to the temperature {tilde over (T)}₀ ⁺used in the charge cycle. Unlike in the configuration in FIGS. 7A and7B, where the working fluid may need to reject a substantial amount ofheat (in sections 38 and 39, respectively) at a fast rate, in thisconfiguration, the hot side and cold side storage media may be cooled atan arbitrarily slow rate (e.g., by radiating away or by other means ofgiving off the heat to the environment).

As shown in FIG. 7D, in some implementations, heat can be rejected fromthe CTS medium to the environment by circulating the CTS medium in thetank 8 in a heat rejection device 55 that can absorb heat from the CTSmedium and reject heat to the environment until the CTS medium coolsfrom the temperature T₀ ⁺ to the temperature {tilde over (T)}₀ ⁺. Insome examples, the heat rejection device 55 can be, for example, aradiator, a thermal bath containing a substance such as water or saltwater, or a device immersed in a natural body of water such as a lake,river or ocean. In some examples, the heat rejection device 55 can alsobe an air cooling apparatus, or a series of pipes which are thermallyconnected to a solid reservoir (e.g., pipes embedded in the ground).

Similarly, in some implementations, heat can be rejected from the HTSmedium to the environment by circulating the HTS in the tank 7 in a heatrejection device 56 that can absorb heat from the HTS medium and rejectheat to the environment until the HTS medium cools from the temperature{tilde over (T)}₁ to the temperature T₁. In some examples, the heatrejection device 56 can be, for example, a radiator, a thermal bathcontaining a substance such as water or salt water, or a device immersedin a natural body of water such as a lake, river or ocean. In someexamples, the heat rejection device 56 can also be an air coolingapparatus or a series of pipes which are thermally connected to a solidreservoir (e.g., pipes embedded in the ground).

In some implementations, rejecting heat to ambient through the use ofthe thermal storage media may be used in conjunction with the variablecompression ratio charge and/or discharge cycles described, for example,in FIG. 5C. In this system, only the CTS medium may exchange heat withambient. Such a system can also be implemented with a recuperator toextend the temperature ranges of the HTS and CTS media in the cycles.

In some implementations, three separate cold side storage tanks atrespective temperatures T₀, {tilde over (T)}₀ ⁺ and T₀ ⁺ may be used(e.g., an extra tank may be used in addition to the tanks 8 and 9).During heat exchange in the cold side CFX 4 in the discharge cycle, heatfrom the working fluid exiting the recuperator 5 may be transferred tothe CTS medium in the T₁ ⁺-tank. The CTS medium may be cooled in/by, forexample, the heat rejection device 55 prior to entering the {tilde over(T)}₀ ⁺-tank. In some implementations, three separate hot side storagetanks at respective temperatures T₁, {tilde over (T)}₁ and T₁ ⁺ may beused (e.g., an extra tank may be used in addition to the tanks 6 and 7).During heat exchange in the hot side CFX 2 in the discharge cycle, heatfrom the working fluid exiting the recuperator 5 may be transferred tothe HTS medium in the {tilde over (T)}₁-tank. The HTS medium may becooled in/by, for example, the heat rejection device 56 prior toentering the T₁-tank. Heat rejection to the environment in such a mannermay present several advantages. In a first example, it may eliminate theneed for a potentially expensive working fluid to ambient heat exchangerthat is capable of absorbing heat from the working fluid at a rateproportional to the power input/output of the system. The HTS and CTSmedia may instead reject heat over extended time periods, thus reducingthe cost of the cooling infrastructure. In a second example, it mayallow the decision regarding when heat is rejected to the environment tobe delayed such that heat exchange to ambient may be performed whentemperature (e.g., the ambient temperature) is most favorable.

In the charge and discharge cycles of FIGS. 7A and 8A, and FIGS. 7B and8B, respectively, the same compression ratios and temperature values areused for both charge and discharge. In this configuration, the roundtripefficiency can be about η_(store)=74%, as given by T₀=194 K (−79° C.),T₁=494 K (221° C.). η_(t)=0.95, η_(c)=0.9 and r=3.3.

Thus, in some examples involving working fluid to working fluidrecuperation, heat rejection on the hot side (high pressure) side of theclosed charge cycle can take place in three operations (heat exchangewith the HTS medium, followed by recuperation, followed by heatrejection to the environment), and heat rejection on the cold side (lowpressure) side of the closed discharge cycle can take place in threeoperations (heat rejection to the environment, followed by recuperation,followed by heat exchange with the CTS medium). As a result ofrecuperation, the higher temperature HTS tank(s) 6 can remain at T₁ ⁺while the lower temperature HTS tank(s) 7 can now be at the temperatureT₁>T₀ ⁺, and the lower temperature CTS tank(s) 9 can remain at T₀ whilethe higher temperature CTS tank(s) 8 can now be at the temperature T₀⁺<T₁.

In some cases, recuperation may be implemented using the heat exchanger5 for direct transfer of heat between the working fluid on the highpressure side and the working fluid on the low pressure side. In analternative configuration, an additional pair (or plurality) of heatexchangers together with an additional heat transfer medium or fluid(e.g., a dedicated thermal heat transfer fluid that is liquid in anappropriate temperature range, such as, for example, therminol) may beused to achieve recuperation. For example, an additional heat exchangermay be added in series with the cold side heat exchanger and anadditional heat exchanger may be added in series with the hot side heatexchanger. The additional heat transfer medium may circulate between thetwo additional heat exchangers in a closed loop. In other examples, oneor more additional heat exchangers may be placed elsewhere in the systemto facilitate recuperation. Further, one or more additional heattransfer media or mixtures thereof may be used. The one or moreadditional heat transfer media fluids may be in fluid or thermalcommunication with one or more other components, such as, for example, acooling tower or a radiator.

In one example, hexane or heptane can be used as a CTS medium, andnitrate salt can be used as an HTS medium. On the low pressure side ofthe cycle, the operating temperatures of the pumped thermal storagecycles may be limited by the melting point of hexane (178 K or −95° C.)at T₀ and by the melting point of the nitrate (494 K or 221° C.) at T₁.On the high pressure side of the cycle, the operating temperatures maybe limited by the boiling point of hexane (341 K or 68° C.) at T₀ ⁺ andby the decomposition of nitrate (873 K or 600° C.) at T₁ ⁺. At theseconditions, the high pressure and low pressure temperature ranges canoverlap such that recuperation can be implemented. The actualtemperatures T₀, T₁, T₀ ⁺ and T₁ ⁺ and pressure ratios implemented inhexane/nitrate systems may differ from the limits above.

In some examples, recuperation may enable the compression ratio to bereduced. In some cases, reducing the compression ratio may result inreduced compressor and turbine losses. In some cases, the compressionratio may be at least about 1.2, at least about 1.5, at least about 2,at least about 2.5, at least about 3, at least about 3.5, at least about4, at least about 4.5, at least about 5, at least about 6, at leastabout 8, at least about 10, at least about 15, at least about 20, atleast about 30, or more.

In some cases, T₀ may be at least about 30 K, at least about 50 K, atleast about 80 K, at least about 100 K, at least about 120 K, at leastabout 140 K, at least about 160 K, at least about 180 K, at least about200 K, at least about 220 K, at least about 240 K, at least about 260 K,or at least about 280 K. In some cases, T₀ ⁺ may be at least about 220K, at least about 240 K, at least about 260 K, at least about 280 K, atleast about 300 K, at least about 320 K, at least about 340 K, at leastabout 360 K, at least about 380 K, at least about 400 K, or more. Insome cases, the temperatures T₀ and T₀ ⁺ can be constrained by theability to reject excess heat to the environment at ambient temperature.In some cases. the temperatures T₀ and T₀ ⁺ can be constrained by theoperating temperatures of the CTS (e.g., a phase transitiontemperature). In some cases, the temperatures T₀ and T₀ ⁺ can beconstrained by the compression ratio being used. Any description of thetemperatures T₀ and/or T₀ ⁺ herein may apply to any system or method ofthe disclosure.

In some cases, T₁ may be at least about 350K, at least about 400 K, atleast about 440 K, at least about 480 K, at least about 520 K, at leastabout 560 K, at least about 600 K, at least about 640 K, at least about680 K, at least about 720 K, at least about 760 K, at least about 800 K,at least about 840 K, at least about 880 K, at least about 920 K, atleast about 960 K, at least about 1000 K, at least about 1100 K, atleast about 1200 K, at least about 1300 K, at least about 1400 K, ormore. In some cases, T₁ ⁺ may be at least about 480 K, at least about520 K, at least about 560 K, at least about 600 K, at least about 640 K,at least about 680 K, at least about 720 K, at least about 760 K, atleast about 800 K, at least about 840 K, at least about 880 K, at leastabout 920 K, at least about 960 K, at least about 1000 K, at least about1100 K, at least about 1200 K, at least about 1300 K, at least about1400 K, at least about 1500 K, at least about 1600 K, at least about1700 K, or more. In some cases. the temperatures T₁ and T₁ ⁺ can beconstrained by the operating temperatures of the HTS. In some cases. thetemperatures T₁ and T₁ ⁺ can be constrained by the thermal limits of themetals and materials being used in the system. For example, aconventional solar salt can have a recommended temperature range ofapproximately 560-840 K. Various system improvements, such as, forexample, increased roundtrip efficiency, increased power and increasedstorage capacity may be realized as available materials, metallurgy andstorage materials improve over time and enable different temperatureranges to be achieved. Any description of the temperatures T₁ and/or T₁⁺ herein may apply to any system or method of the disclosure.

In some cases, the roundtrip efficiency η_(store) (e.g., electricitystorage efficiency) with and/or without recuperation can be at leastabout 5%, at least about 10%, at least about 15%, at least about 20%, atleast about 25%, at least about 30%, at least about 35%, at least about40%, at least about 45%, at least about 50%, at least about 55%, atleast about 60%, at least about 65%, at least about 70%, at least about75%, at least about 80%, at least about 85%, at least about 90%, or atleast about 95%.

In some implementations, at least a portion of heat transfer in thesystem (e.g., heat transfer to and from the working fluid) during acharge and/or discharge cycle includes heat transfer with theenvironment (e.g., heat transfer in sections 38 and 39). The remainderof the heat transfer in the system can occur through thermalcommunication with thermal storage media (e.g., thermal storage media 21and 22), through heat transfer in the recuperator 5 and/or throughvarious heat transfer processes within system boundaries (i.e., not withthe surrounding environment). In some examples, the environment mayrefer to gaseous or liquid reservoirs surrounding the system (e.g., air,water), any system or medium capable of exchanging thermal energy withthe system (e.g., another thermodynamic cycle or system, heating/coolingsystems, etc.), or any combination thereof. In some examples, heattransferred through thermal communication with the heat storage mediacan be at least about 25%, at least about 50%, at least about 60%, atleast about 70%, at least about 80%, or at least about 90% of all heattransferred in the system. In some examples, heat transferred throughheat transfer in the recuperator can be at least about 5%, at leastabout 10%, at least about 15%, at least about 20%, at least about 25%,at least about 50%, or at least about 75% of all heat transferred in thesystem. In some examples, heat transferred through thermal communicationwith the heat storage media and through heat transfer in the recuperatorcan be at least about 25%, at least about 50%, at least about 60%, atleast about 70%, at least about 80%, at least about 90%, or even about100% of all heat transferred in the system. In some examples, heattransferred through heat transfer with the environment can be less thanabout 5%, less than about 10%, less than about 15%, less than about 20%,less than about 30%, less than about 40%, less than about 50%, less thanabout 60%, less than about 70%, less than about 80%, less than about90%, less than about 100%, or even 100% of all heat transferred in thesystem. In some implementations, all heat transfer in the system may bewith the thermal storage media (e.g., the CTS and HTS media), and onlythe thermal storage media may conduct heat transfer with theenvironment.

Pumped thermal cycles of the disclosure (e.g., the cycles in FIGS. 7Aand 7B) may be implemented through various configurations of pipes andvalves for transporting the working fluid between the turbomachinery andthe heat exchangers. In some implementations, a valving system may beused such that the different cycles of the system can be interchangedwhile maintaining the same or nearly the same temperature profile acrossat least one, across a subset or across all of counter-flow heatexchangers in the system. For example, the valving may be configuredsuch that the working fluid can pass through the heat exchangers inopposite flow directions on charge and discharge and flow directions ofthe HTS and CTS media are reversed by reversing the direction of thepumps.

In some implementations, the system may be set up to enable switchingbetween different cycles, such as, for example, the cycles in FIGS. 7Cand 7D, using a shared set of valves and pipes. A schematic of howpiping and valves may be arranged to achieve working fluid flow in acounter-flow thermodynamic cycle is shown in FIG. 7E for charge (such asthe cycle shown, for example, in FIG. 7C) and in FIG. 7F for discharge(such as the cycle shown, for example, in FIG. 7D). The counter-flowheat exchangers can include the hot side CFX 2, the cold side CFX 4and/or the recuperator 5. Such a configuration may be advantageous as itmay reuse at least a portion, or a substantial portion or majority ofthe same piping for the working fluid in both the charging anddischarging modes. While the working fluid changes may change directionbetween charge and discharge, the temperature profile of the heatexchangers can be kept constant, partially constant, or substantially orfully constant by changing the direction in which the HTS medium 68 andthe CTS medium 69 are pumped when switching from charge to discharge andvice versa, and/or by matching the heat fluxes of the working fluid, theHTS medium 68 and the CTS medium 69 appropriately.

Symbols 65 represent either three-way or four-way valves. The whitequadrants indicate the direction in which the working fluid is allowedto pass, black quadrants indicate where flow is blocked. For example, avalve symbol with white quadrants on left and right and black quadrantson top and bottom will allow the working fluid to pass from left toright and not in any other direction. The paths 66 represent a series ofpipes in which the working fluid 20 flows. The pipes 66 may in somecases be thermally insulated and/or rated at a given pressure or rangeof pressures. The pipes 66 may be so designed as to minimize losses dueto turbulence and or pressure drop. Working fluid flow is indicated bylight arrows inside the pipes 66.

The examples in FIGS. 7C and 7D illustrate an operating principle of avalving and piping system configured such that the different cycles ofthe system can be interchanged while maintaining the same, partially thesame, or substantially the same temperature profile across at least oneof the counter-flow heat exchangers, across a subset of the counter-flowheat exchangers, or across all of the counter-flow heat exchangers inthe system. This may be advantageous as the heat capacities of the heatexchangers themselves can mean that changing their temperature profilebetween cycles could result in additional system losses. The exampleschematics also demonstrate the design principle of striving to keep thetotal amount of piping and valving to a minimum by reusing them fordifferent cycles where possible. In various examples, differentconfigurations of piping and valves, and/or a different number of pipesor valves may be used.

Solar Assisted Pumped Thermal Storage Cycles

Another aspect of the disclosure relates to pumped thermal systemsassisted by solar heating. Since the pumped thermal systems of thedisclosure can be configured to store electricity using heat, thesystems may be enhanced by utilizing other sources of heat (e.g.,concentrating solar heat, waste heat, combustion, etc.). In some cases,the discharge cycle can be the same as for systems without solarheating. In some cases, the discharge cycle can be different than forsystems without solar heating. In some cases, a separate charge cyclecan be used on the hot side of the system for recharging the HTS mediumand/or a separate charge cycle can be used on the cold side of thesystem for recharging the CTS medium. In some examples, a combinedcharge cycle for the hot and cold sides can be used.

In an example, in a discharge mode, a solar assisted pumped thermalstorage system can operate on a discharge cycle such as, for example,the discharge cycle in FIG. 7B or FIG. 7D. In a charge mode, the heatfor affecting a temperature increase of the HTS medium from T₁ to T₁ ⁺can be at least partially supplied by solar heating instead of by heatexchange with the working gas on the hot side of the system as in, forexample, the charge cycle in FIG. 7A. This may recharge the HTS mediumon the hot side, but not the CTS medium on the cold side. The CTS mediumcan be recharged separately by using a refrigerator cycle. In anotherexample, the charge cycle in FIG. 7A can be used in concert with solarheating.

FIG. 9A is a schematic flow diagram of cold side recharging in a pumpedheat cycle that can be coupled with external (e.g. solar) heat input.The working fluid 20 enters the compressor at T₁ and P₂, exits thecompressor at T₁ ⁺ and P₁, rejects heat Q₁ to an intermediate thermalstorage (ITS) medium 23 in the hot side CFX 2, exiting the hot side CFX2 at {tilde over (T)}₀ ⁺ and P₁, enters the turbine 3 at {tilde over(T)}₀ ⁺ and P₁, exits the turbine at T₀ and P₂, absorbs heat Q₂ from theCTS medium 22 in the cold side CFX 4, and exits the cold side CFX 4 atT₁ and P₂, returning to its initial state before entering thecompressor. The temperature T₁ ⁺ of the working fluid exiting thecompressor and entering the hot side CFX 2 is higher than thetemperature of the ITS medium 23 entering the hot side CFX 2 from asecond intermediate thermal storage tank 15 at a temperature {tilde over(T)}₀ ⁺. Thus, the working fluid exits the hot side CFX 2 at thetemperature {tilde over (T)}₀ ⁺ and the ITS medium exits the hot sideCFX 2 into a first intermediate thermal storage tank 14 at thetemperature T₁ ⁺. The heat exchange process can take place at a constantor near-constant pressure such that the working fluid exits the hot sideCFX 2 at a lower temperature but same pressure P₁. Similarly, thetemperature of the ITS medium 23 increases in the hot side CFX 2, whileits pressure can remain constant or near-constant. The two intermediatethermal storage tanks 14 and 15 on the hot side can have the samefunctionality as the hot side thermal storage tanks 6 and 7; however,their temperature is lower for reasons described next. The ITS medium(or media) in this configuration may be the same as one or more HTSmedia described herein. Alternatively, the ITS medium may partially orfully distinct from HTS media described herein. In one example, the ITSmedium can include one or more additional heat storage media. In anotherexample, the ITS medium can include the same components as an HTSmedium, but in different proportions (e.g., different mixture). In someimplementations, the ITS medium 23 may be a liquid over the entire rangefrom {tilde over (T)}₀ ⁺ to T₁ ⁺. In other implementations, the ITSmedium 23 may be not be a liquid over the entire range from {tilde over(T)}₀ ⁺ to T₁ ⁺ but may be provided to the counter-flow heat exchanger 2at a higher flow rate in order to achieve a lower temperature riseacross the counter-flow heat exchanger (e.g., such that the temperatureof the ITS at the exit of the counter-flow heat exchanger 2 is lowerthan T₁ ⁺) while still cooling the working fluid from T₁ ⁺ to {tildeover (T)}₀ ⁺. In this instance, the temperature of the ITS medium in thetank 14 can be lower than T₁ ⁺. The ITS medium in the tank 14 canexchange heat with ambient (e.g., through a radiator or otherimplementations described herein) to be cooled down back to thetemperature T₀ ⁺. The ITS medium can then be returned back into the tank15. The heat deposited in the ITS medium may be used for various usefulpurposes, such as, for example, residential or commercial heating,thermal desalination or the other uses mentioned elsewhere herein. Insome implementations, the lower temperature {tilde over (T)}₀ ⁺ of theITS can be at about ambient temperature. Depending on the climate of thesurrounding region, the ambient temperature can vary significantly, andsystem operation can be varied accordingly. In some implementations, thelower temperature {tilde over (T)}₀ ⁺ of the ITS medium can be at atemperature higher than the ambient temperature.

In some implementations, the recharge cycle in FIG. 9A can be modifiedto include heat rejection to the environment (or other heat sink)equivalent to heat rejection in section 38 (e.g., a radiator) describedelsewhere herein. In such a case, the temperature of the ITS medium inthe tank 15 may be T₀ ⁺ instead of {tilde over (T)}₀ ⁺.

FIG. 10 shows a cold side recharge cycle for a hexane/salt system insolar mode in accordance with the cold side recharge cycle in FIG. 9Awith η_(c)=0.9 and η_(t)=0.95, modified to include heat rejection to theenvironment in section 38. An objective of the cold side recharge cyclemay be to refrigerate the CTS medium while expending as little work aspossible. In some examples, this can involve using the lowest possiblecompression ratio while maintaining a workable refrigerator cycle. Insome cases, a lower limit of the compression ratio can be constrained tovalues for which the temperature T₀ ⁺ remains sufficiently high withrespect to ambient temperature to ensure that heat can be still berejected to the environment between the temperatures T₀ ⁺ and {tildeover (T)}₀ ⁺ as indicated in FIG. 10. In one example, the compressor 3of the cold side recharge cycle can operate between {tilde over (T)}₀⁺=300 K (27° C.) and T₀=189 (−84° C.), corresponding to a compressionratio r=3.38. Thus, in examples where lower pressure ratios are used inthe cold side recharge cycles, the temperatures of the thermal storagetanks 14 and 15 can be lower than the temperatures of thermal storagetanks with the same functionality operating in cycles with higherpressure ratios, such as, for example, in the charge cycles in FIGS. 7Aand 8A.

Further, lower values of the temperature T₁ may be used in the cold siderecharge cycles. In some examples using separate hot side recharging,the cold side recharge cycles can be configured for recharging of thecold side only. For example, the cold side recharge system and themodified cycle in FIGS. 9A and 10 can operate at T₀=189 K (−84° C.),T₁=350.7 K (78° C.) and r=3.38. Lower temperatures T₁ ⁺ and T₀ ⁺ may beused than if the system in FIG. 9A were also used for recharging the HTSmedium 21. In some implementations, the recuperator 5 (e.g., as used inthe corresponding discharge cycle in FIG. 7B) can be bypassed in part orcompletely during the recharging of the cold side, T₁ can be at or nearambient temperature, resulting in the compression of colder workingfluid compared to the charge cycle shown, for example, in FIG. 7A,leading to less work W₁ consumed by the compressor 1 to increase thepressure and temperature of the working fluid to T₁ ⁺, P₁. In otherimplementations, partially or completely separate systems for cold siderecharging and hot side recharging may be used. For example, the hotside recharging system and cold side recharging systems can have 0, 1,2, or more joint system components (e.g., turbomachinery).

FIG. 9B is another example of cold side recharging in a pumped heatcycle that can be coupled with external (e.g. solar) heat input. In thisexample, the temperature of one of the two CTS tanks is higher thanambient temperature. In this configuration, improved efficiency may beachieved by allowing the CTS medium 22 in the thermal storage tank 8having a temperature {tilde over (T)}₁ to exchange heat with thesurroundings, thus cooling the CTS medium to a temperature T₁, equal toabout the ambient temperature, before using the refrigerator (cold siderecharge cycle) to further cool the CTS medium from T₁ to T₀. In thisconfiguration, the system can be operated at, for example, T₀=189 K(−84° C.), T₁=300 K (27° C.), and r=3.38. In some cases, the ambientcooling in FIG. 9B may be particularly suited for operation at night,when the ambient temperature can be lower.

FIG. 9F shows another example of cold side recharging in which heat isrejected to the ambient environment. An intermediate medium or fluid 57(e.g., therminol, or a heat transfer oil) that is liquid over a suitablerange of operating temperatures may be used for exchanging heat betweenthe working fluid 20 and a thermal bath 58 in the hot side CFX 2. Theuse of the intermediate fluid 57 may provide an advantage overcontacting an inexpensive thermal sink or medium (e.g., water) directlywith the working fluid. For example, directly contacting such a thermalmedium with the working fluid in the hot side CFX 2 may cause problems,such as, for example, evaporation or over-pressurization (e.g.,explosion) of the thermal medium. The intermediate fluid 57 can remainin liquid phase throughout all, a portion of, or a significant portionof the operation in the hot side CFX 2. As the intermediate fluid 57passes through the thermal bath 58, it can be sufficiently cooled tocirculate back into the hot side CFX 2 for cooling the working fluid(e.g., from T₁ ⁺ to {tilde over (T)}₀ ⁺). The thermal bath 58 maycontain a large amount of inexpensive heat sink material or medium, suchas, for example, water. In some cases, the heat deposited in this heatsink may be used for useful purposes such as, for example, residentialor commercial heating, thermal desalination or the other uses describedelsewhere herein. In some cases, the heat sink material may bere-equilibrated with ambient temperature (e.g., through a radiator orother implementations described herein).

In some implementations, the cold side recharge cycles in FIGS. 9A, 9Band/or 9F may include a recuperator, as described in greater detail inexamples throughout the disclosure. Such systems may be implementedusing the temperatures T₁ ⁺, T₁, T₀ ⁺ and T₀ described in greater detailelsewhere herein.

FIG. 9C is a schematic flow diagram of hot side recharging in a pumpedheat cycle in solar mode with heating of a solar salt solely by solarpower. The system can comprise a solar heater for heating the hot sideheat storage. The HTS medium 21 in the second hot thermal storage tank 7of a discharge cycle, such as, for example, the HTS medium of thedischarge cycle in FIG. 7B, can be recharged within element 17 usingheating provided by solar radiation. The HTS medium (e.g., molten salt)can be heated by solar heating from the temperature T₁ in the second hotthermal storage tank 7 to the temperature T₁ ⁺ in the first hot thermalstorage tank 6.

FIG. 9D is a schematic flow diagram of hot side recharging in a pumpedheat cycle in solar mode using a heat exchanger between an intermediatefluid tank and a solar salt. The HTS medium 21 (e.g., a molten salt orsolar salt) in the second hot thermal storage tank 7 of a dischargecycle, such as, for example, the HTS medium at about 493 K (220° C.) inthe discharge cycle in FIG. 7B, can be recharged by first exchangingheat in a heat exchanger 18 with a hotter thermal storage medium, suchas, for example, the ITS medium 23 at about 595 K (322° C.) in the firstintermediate thermal storage tank 14 in FIG. 9A, followed by solarheating within element 17 to about 873 K (600° C.). The HTS medium(e.g., molten salt) can be heated from the temperature T₁ in the secondhot thermal storage tank 7 to the temperature T₁ ⁺ in the first hotthermal storage tank 6. The hotter thermal storage medium 23 may be theITS medium in FIG. 9A, or any other thermal storage medium or waste heatstream at a suitable temperature. After exchanging heat with the HTSmedium 21, the thermal storage medium can be stored in a thirdintermediate thermal storage tank 16 and/or can be used as a heatexchange fluid in the systems herein, for cogeneration etc.

In some implementations, such as, for example, for the systems in FIGS.9C and/or 9D, solar heat for heating the HTS medium (e.g., from T₁=493 K(220° C.) to T₁ ⁺=873 K (600° C.)) may be provided by a concentratingsolar facility. In some examples, a small scale concentrating facilitymay be utilized for providing heat. In some cases, the concentratingsolar facility may include one or more components for achieving highsolar concentrating efficiency, including, for example, high-performanceactuators (e.g., adaptive fluidic actuators manufactured from polymers),mutiplexing control system, dense heliostat layout etc. In someexamples, the heat provided for heating the HTS medium (e.g., in theelement 17) may be a waste heat stream from the concentrating solarfacility.

FIG. 9E is a schematic flow diagram of a pumped thermal system chargecycle with a gas-gas heat exchanger in parallel with solar heat input.In this configuration, the pumped thermal charge mode in FIG. 7A (usingthe gas-gas heat exchanger 5 and a combined charge cycle for both hotand cold side recharging) is augmented by parallel solar heating in theelement 17. The charge cycle can have two parallel means of rechargingHTS (e.g., molten salt) tanks. In some implementations, electricityproduced by one or more power sources (e.g., a photovoltaic power plant)during the day (e.g., when ambient temperatures are high) can be storedusing the systems of FIG. 9E. Solar heating may be implementedsimultaneously for charging the hot side storage tanks containing HTSmedium (e.g., molten salt). In one example, separate molten salt pipesand pumps can be used for the electricity storage cycle and for thesolar heating. In another example, one or more components can be shared,such as, for example, molten salt pipes, pumps, one or more hot sidestorage tanks etc. In some cases, the element 17 can be a solarconcentrating heating element (e.g., a channel in which a molten salt isflowing and onto which solar radiation is concentrated). An excess of“charge” in the hot side tanks compared to the cold side tanks can beachieved (e.g., a high fill level of the HTS medium in the first hotthermal storage tank 6 relative to the CTS medium in the second coldthermal storage tank 9). The cold side tanks not recharged by the chargecycle in FIG. 9E may be recharged using the systems of FIGS. 9A and/or9B. For example, the cold side tanks can be recharged at night when theambient temperature is low and cooling (e.g., ambient cooling the heatexchanger 27 in FIG. 9B) is easier to accomplish. In another example,the cold side tanks can be recharged at any time using, for example, thesystem in FIG. 9A, operating independently of the charge cycle in FIG.9E (or any other electricity storage charge cycle of the disclosure).

In some implementations, cold side, hot side and/or combined chargecycles herein can be used with the discharge cycle in FIGS. 7B and 8B toachieve various levels of roundtrip efficiencies. In some cases,addition of solar heat can result in significant electricity storageroundtrip efficiency improvements. In some cases, the roundtripefficiency η_(store) (e.g., electricity storage efficiency) withoutand/or with solar heat addition can be at least about 25%, at leastabout 50%, at least about 75%, at least about 100%, at least about 125%,at least about 150%, at least about 175%, at least about 200%, at leastabout 225%, at least about 250%, at least about 300%, at least about350%, at least about 400%, at or more.

In one example, the cold side recharge cycle in FIGS. 9A and 10 can beused with the hot side recharge cycle in FIG. 9C and the discharge cyclein FIGS. 7B and 8B. In this configuration, the compression ratios in thecharge and discharge cycles can be different (e.g., for the charge cyclein FIG. 9A, T₀=215 K (−57° C.), T₁=300 K (27° C.) and r=2.45, while forthe discharge cycle in FIG. 7B, T₀=215 K (−57° C.), T₁=563 K (290° C.)and r=2.95), with a roundtrip storage efficiency (also “roundtripefficiency” herein) of η_(store)=227% and a value of T₁ ⁺ upon cold siderecharge of 447 K (173° C.). In some cases, at least a portion of theheat stored in the ITS medium in the first intermediate thermal storagetank 14 in the cold side recharge cycle in FIGS. 9A and 10 can be usedto charge the HTS medium (e.g., molten salt) used in the discharge cyclein FIGS. 7B and 8B using, for example, the system in FIG. 9D to reducethe amount of solar concentration needed compared to the system in FIG.9C. For example, the HTS medium in the discharge cycle can have a lowtemperature of about 493 K (220° C.), while the ITS medium in therefrigerator cycle can have a temperature of up to about 593 K (320°C.). Instead of raising the temperature of the HTS medium (e.g., moltensalt) using solar heat from 493 K (220° C.) to 593 K (320° C.), it maybe raised by first exchanging heat with the ITS medium, and only thenusing solar heat to heat the HTS medium the rest of the way.

In another example, where the temperature of the CTS medium in the tank8 is at a temperature {tilde over (T)}₁ greater than ambienttemperature, the cold side recharge cycle in FIG. 9B can be used withthe hot side recharge cycle in FIG. 9C and the discharge cycle in FIGS.7B and 8B. In this configuration, the compression ratios in the chargeand discharge cycles can again be different.

In another example, in cases where it may be difficult to vary thecompression ratio between charge (e.g., cold side recharge) anddischarge, both the cold side recharge cycle compression ratio and thedischarge cycle compression ratio can be set to r=2.45. In this example,for a system operating according to the charge cycle in FIG. 9A, T₀=215K (−57° C.), T₁=300 K (27° C.) and r=2.45, while for the discharge cyclein FIG. 7B, T₀=215 K (−57° C.), T₁=563 K (290° C.) and r=2.45, aroundtrip storage efficiency (also “roundtrip efficiency” herein) ofη_(store)=208% and a value of T₁ ⁺ upon cold side recharge of 447 K(173° C.) may be achieved.

FIG. 9G is a schematic flow diagram of a pumped thermal system dischargecycle that can be coupled with external heat input (e.g., solar,combustion) with heat rejection to ambient. Such a discharge cycle maybe used, for example, in situations where the capacity for hot siderecharging (e.g., using solar heating, waste heat or combustion) isgreater than the capacity for cold side recharging. Solar heat may beused to charge the HTS medium 21 in the hot side storage tanks from T₁to T₁ ⁺, as described elsewhere herein. The discharge cycle can operatesimilarly to the discharge cycle in FIG. 2B, but after exiting theturbine 3, the working fluid 20 can proceed to the cold side CFX 4 heatexchanger 4 where it exchanges heat with an intermediate thermal storage(ITS) medium 61 having a lower temperature T₀ at or near ambienttemperature. The ITS medium 61 enters the cold side CFX 4 from a secondintermediate thermal storage tank 59 at the temperature T₀ (e.g.,ambient temperature) and exits the cold side CFX 4 into a firstintermediate thermal storage tank 60 at the temperature {tilde over(T)}₁, while the working fluid 20 enters the cold side CFX 4 at thetemperature {tilde over (T)}₁ and exits the cold side CFX 4 at thetemperature T₀. The working fluid enters the compressor 1 at T₀ and P₂,exits the compressor at T₀ ⁺ and P₁, absorbs heat Q₁ from the HTS medium21 in the hot side CFX 2, exits the hot side CFX 2 at T₁ ⁺ and P₁,enters the turbine 3 at T₁ ⁺ and P₁, exits the turbine at {tilde over(T)}₁ and P₂, rejects heat Q₂ from the ITS medium 61 in the cold sideCFX 4, and exits the cold side CFX 4 at T₀ and P₂, returning to itsinitial state before entering the compressor.

In some implementations, the ITS medium 61 may be a liquid over theentire range from T₀ to {tilde over (T)}₁. In other implementations, theITS medium 61 may not be a liquid over the entire range from T₀ to{tilde over (T)}₁, but may be provided to the counter-flow heatexchanger 4 at a higher flow rate in order to achieve a lowertemperature rise across the counter-flow heat exchanger (e.g., such thatthe temperature of the ITS medium at the exit of the counter-flow heatexchanger 4 is lower than {tilde over (T)}₁) while still cooling theworking fluid from {tilde over (T)}₁ to T₀. In this instance, thetemperature of the ITS medium in the tank 60 can be lower than {tildeover (T)}₁. The ITS medium in the tank 60 can exchange heat with ambient(e.g., through a radiator or other implementations described herein) inorder to cool back to the temperature T₀. In some cases, the ITS mediumcan then be returned to the tank 59. In some cases, the heat depositedin the ITS medium may be used for various useful purposes such as, forexample, residential or commercial heating, thermal desalination orother uses described elsewhere herein.

FIG. 9H is a schematic flow diagram of a pumped thermal system dischargecycle in solar mode with heat rejection to an intermediate fluidcirculated in a thermal bath at ambient temperature. The discharge cyclecan operate similarly to the discharge cycle in FIG. 9G, but afterexiting the turbine 3, the working fluid 20 can proceed to the cold sideCFX 4 where it exchanges heat with an intermediate medium or fluid 62circulating through a thermal bath 63 at the temperature T₀ at or nearambient temperature. The intermediate medium or fluid 62 (e.g.,therminol, or a heat transfer oil) may be used for exchanging heatbetween the working fluid 20 and a thermal bath 63 in the cold side CFX4. The use of the intermediate fluid 62 may provide an advantage overcontacting an inexpensive thermal sink or medium (e.g., water) directlywith the working fluid. For example, directly contacting such a thermalmedium with the working fluid in the cold side CFX 4 may cause problems,such as, for example, evaporation or over-pressurization (e.g.,explosion) of the thermal medium. The intermediate fluid 62 can remainin liquid phase throughout all, at least a portion of, or a significantportion of the operation in the cold side CFX 4. As the intermediatefluid 62 passes through the thermal bath 58, it can be sufficientlycooled to circulate back into the cold side CFX 4 for cooling theworking fluid from {tilde over (T)}₁ to T₀. The thermal bath 63 maycontain a large amount of inexpensive heat sink material or medium, suchas, for example, water. In some cases, the heat deposited in the heatsink material may be used for various useful purposes such as, forexample, residential or commercial heating, thermal desalination orother uses described elsewhere herein. In some cases, the heat sinkmaterial may be re-equilibrated with ambient temperature (e.g., througha radiator or other implementations described herein).

In some implementations, the discharge cycles in FIGS. 9G and/or 9H mayinclude a recuperator, as described in greater detail in examplesthroughout the disclosure. Such systems may be implemented using thetemperatures T₁ ⁺, T₁, T₀ ⁺ and T₀ described in greater detail elsewhereherein.

Addition of solar heat may lead to more work being extracted than workoriginally supplied to the system via electrical energy. The roundtripefficiency may thus be larger than 100%. In some examples, thermalefficiency for solar assisted systems can be expressed asη_(thermal)=W_(extra)/Q_(solar), where W_(extra) is the extra workoutput from a charge/discharge cycle with solar heat input Q_(solar) ascompared to the work output from a charge/discharge cycle without solarheat input. In some examples, heat from cold side charge cycles can beutilized, as shown, for example, in FIG. 9D, to lower the Q_(solar)needed, thus raising the thermal efficiency (thermal of the solarassisted pumped thermal storage cycles herein.

In some examples, the thermal efficiency η_(thermal) (e.g., heat toelectricity conversion efficiency) of solar assisted pumped thermalstorage cycles can be at least about 5%, at least about 10%, at leastabout 15%, at least about 20%, at least about 25%, at least about 30%,at least about 35%, at least about 40%, at least about 45%, at leastabout 50%, at least about 60%, at least about 70%, or more.

An additional way of characterizing the performance of storage systemsthat take both heat and electricity as inputs and provide electricity asan output can be referred to as an exergetic efficiency of the system.Exergy can be defined as the potential for a given amount of energy tobe converted into useful work. Electrical energy may be approximated tobe equal to exergy since it can be converted to work with minimallosses. Thermal energy, however, can be converted to work by bringing ahigher temperature thermal bath into equilibrium with a lowertemperature thermal bath. A limit of how much work can be converted froma given amount of heat or thermal energy Q at a temperature T_(H) cominginto equilibrium with a lower temperature bath T_(C) can be given byCarnot efficiency,

$W_{carnot} \leq {{Q\left( {1 - \frac{T_{C}}{T_{H}}} \right)}.}$

the quantity W_(carnot) can also be called a thermal exergy contentX_(Th) of the heat Q. For heat added over a range of temperatures, forexample from T₁ to T₁ ⁺, the thermal exergy content for constantspecific heat c_(p) can be

$X_{Th} = {{\int_{T_{1}}^{T_{1}^{+}}{dQ}} = {T_{1}^{+} - T_{1} - {T_{C}{{\ln \left( \frac{T_{1}^{+}}{T_{1}} \right)}.}}}}$

For a system that takes in electricity exergy X_(elec) ^(in) and thermalexergy X_(Th), and outputs electricity exergy X_(elec) ^(out), theexergetic efficiency can be defined as

$\eta_{x} = {\frac{X_{elec}^{out}}{X_{elec}^{in} + X_{Th}}.}$

in some examples, systems of the disclosure may have an exergeticefficiency η_(x) of at least about 5%, at least about 10%, at leastabout 15%, at least about 20%, at least about 25%, at least about 30%,at least about 35%, at least about 40%, at least about 45%, at leastabout 50%, at least about 60%, at least about 70%, at least about 80%,or at least about 90%, or more.

Solar Assisted Pumped Thermal Storage Cycles with Intercooling

An aspect of the disclosure provides intercooling of pumped thermalstorage systems. In some implementations, intercooling of solar assistedpumped thermal storage systems is provided. In an example, a cold siderecharge (refrigerator) cycle in a solar assisted pumped thermal storagesystem can be aimed at cooling cold side thermal storage tanks, withlimited utility of heating the working fluid in the refrigerator cycle.The hotter the working fluid, the more work may be required to compressit. In some cases, at least a portion of the sensible energy of theworking fluid may be used for preheating an HTS medium (e.g., solarsalt), for example, as shown in FIG. 9D, via direct heat exchangebetween the working fluid and the HTS medium, or any combination orvariation thereof. Further, in some cases, at least a portion of thesensible energy of the working fluid and/or the ITS medium of the coldside recharge cycle may transferred to another medium, such as, forexample, a process or residential heating stream. As an alternative, incases where heat transfer is not employed, compression work (e.g., workinput W₁ in FIGS. 9A and 9B) may be decreased or minimized byapproaching isothermal operation of the compressor 1. Examples includeachieving a temperature rise (T₁ ⁺−T₁) in the compressor 1 of less thanabout 2% of the compressor inlet temperature T₁, less than about 5% ofthe compressor inlet temperature T₁, less than about 10% of thecompressor inlet temperature T₁, less than about 15% of the compressorinlet temperature T₁, less than about 20% of the compressor inlettemperature T₁, less than about 25% of the compressor inlet temperatureT₁, less than about 50% of the compressor inlet temperature T₁, lessthan about 75% of the compressor inlet temperature T₁ less than about95% of the compressor inlet temperature T₁, less than about 120% of thecompressor inlet temperature T₁, or less than about 150% of thecompressor inlet temperature T₁, or more. In an example, substantiallyisothermal compression may be achieved.

In one example, work input for compressing one mole of an ideal gas inan ideal (e.g., reversible) compressor operating isothermally (also“isothermal limit” herein) can be expressed as

${W_{{isot}.{comp}}^{ideal}} = {{\overset{\_}{R}T{\int_{V_{1}}^{V_{2}}\frac{dV}{V}}} = {{\overset{\_}{R}\ln \frac{V_{2}}{V_{1}}} = {{\overset{\_}{R}\ln \frac{p_{1}}{p_{2}}} = {\left( {{\overset{\_}{c}}_{p}\frac{\gamma_{- 1}}{\gamma}} \right)T_{inlet}{{\ln (r)}.}}}}}$

In practical systems using turbomachinery, isothermal compression may beapproximated, or in some cases, partially or substantially achieved, byutilizing intercooling to reduce the temperature of the working fluidbetween compression stages. In some cases, each compression stage may bemodeled as an adiabatic compression stage having a given adiabatic(isentropic) efficiency.

FIGS. 11A-D show (clockwise) a solar assisted cold side recharge cyclewith intercooling with 0, 1, 2 and 3 stages, respectively. In FIG. 11A(no intercooling), the working fluid enters the compressor 1 at T₁ andP₂, exits the compressor at T₁ ⁺ and P₁, rejects heat in the hot sideCFX 2, exits the hot side CFX 2 at T₀ ⁺ and P₁, rejects heat to theenvironment (or other heat sink) in section 38 (e.g., a radiator),enters the turbine 3 at {tilde over (T)}₀ ⁺ and P₁, exits the turbine atT₀ and P₂, absorbs heat in the cold side CFX 4, and exits the cold sideCFX 4 at T₁ and P₂, returning to its initial state before entering thecompressor. In FIG. 11B (1 intercooling stage), the working fluid entersthe compressor 1 at T₁ and P₂, but exits the compressor at a temperatureT⁺ _(1,int) and P₁, where T₁<<T⁺ _(1,int). In FIG. 11B (2 intercoolingstages), the working fluid exits the compressor at a temperature T₁<T⁺_(1,int)<T₁ ⁺ and P₁, but with T⁺ _(1,int) now being lower than in FIG.11B. Similarly, an even lower temperature T⁺ _(1,int) can be achievedwith 3 intercooling stages FIG. 11C. With additional intercooling stagesadded, the temperature at the compressor exit may approach thecompressor inlet temperature T₁ (i.e., approaches an isothermal process)and the required compression work may be decreased.

FIG. 12 illustrates the effect of additional stages of intercooling oncompression work W_(comp) for a cold side recharge cycle in a solarassisted pumped thermal storage system with r=3.38 and T₁=300 K (27°C.). In some examples, intercooling can be implemented with equalcompression ratios (and, in some cases, equal temperature rises) at eachstage. In this configuration, intercooling stages can occur at pressures

${p_{2}\left( \frac{p_{1}}{p_{2}} \right)}^{\frac{1}{N}},{{p_{2}\left( \frac{p_{1}}{p_{2}} \right)}^{\frac{2}{N}}\mspace{14mu} \ldots \mspace{14mu} {p_{2}\left( \frac{p_{1}}{p_{2}} \right)}^{\frac{N - 1}{N}}}$

and work done for N_(int) stages of intercooling can be given by, forexample,

${W_{comp} = {{\overset{\_}{c}}_{p}T_{inlet}{N_{int}\left( {\psi^{\frac{1}{N_{int}\eta_{cp}}} - 1} \right)}}},$

where, as N_(int)→∞, W_(comp) approaches the isothermal limit, as shownin FIG. 12, where the isothermal limit is represented by a horizontalline. In some examples, compressors with N_(int) stages of intercoolingcan be utilized in solar assisted pumped thermal storage systems.

Intercooling of one or more compressors (e.g., during cold siderecharge), reheating of one or more turbines (e.g., during discharge),recuperation, heat exchange between two or more cycles (e.g., as in, forexample, FIG. 9D), utilization of external sources of heat (e.g., solarheat for recharging hot side thermal storage tanks), cold (e.g., coldenvironment or liquefaction/cryogenic reservoirs for recharging coldside thermal storage tanks) and/or waste heat/cold (e.g., industrialprocess heat or cold for preheating, precooling, reheating, intercoolingetc.), and/or other strategies for improving efficiency may be employedin pumped thermal storage systems of the disclosure. In someimplementations, flow rates of thermal storage media and/or the workingfluid may be adjusted to accommodate the incorporation of additionalheat sources and/or heat sinks.

In some instances, the pumped thermal system may provide heat sourcesand/or cold sources to other facilities or systems such as, for example,through co-location with a gas to liquids (GTL) facility or adesalination facility. In one example, the GTL facilities may make useof one or more of the cold reservoirs in the system (e.g., the CTSmedium in the tank 9 for use in oxygen separation in the GTL facility)and/or one or more hot reservoirs in the system (e.g., the HTS medium inthe tank 6 for use in a Fischer-Tropsch process in the GTL facility). Inanother example, one or more hot reservoirs or one or more coldreservoirs in the pumped thermal system may be used for the operation ofthermal desalination methods. Further examples of possible heat and colduses include co-location or heat exchange with building/area heating andcooling systems.

Conversely, in some cases, the pumped thermal system may make use ofwaste heat sources and/or waste cold sources from other facilities orsystems such as, for example, through co-location with a liquefiednatural gas import or export terminal. For example, a waste cold sourcemay be used for cooling the cold side thermal storage media 22. In someimplementations, recharging of the cold side using waste cold may becombined with recharging of the hot side thermal storage media 21 byexternal heat input (e.g., solar, combustion, waste heat, etc.). In somecases, the recharged storage media can then be used in a discharge cyclesuch as, for example, the discharge cycles in FIG. 7B or 7D. In somecases, the pumped thermal system may be used as a heat engine with awaste heat source serving as the hot side heat input and a waste coldsource serving as the cold side heat sink. In another implementation,the hot side storage media may be recharged using a modified version ofthe cycle shown in FIG. 7C, where the temperature T₀ is about theambient temperature and {tilde over (T)}₀ ⁺ corresponds to a temperatureabove the ambient temperature. In some examples, a waste heat source canbe used to provide the heat needed at a temperature of at least {tildeover (T)}₀ ⁺ for heating the working fluid and/or the CTS medium to{tilde over (T)}₀ ⁺. In another implementation, an intermediate fluid(e.g., therminol) which can remain liquid between the temperatures{tilde over (T)}₀ ⁺ and T₀ may be used to transfer the heat from thewaste heat source to the working fluid.

Pumped Thermal Systems with Dedicated Compressor/Turbine Pairs

In a further aspect of the disclosure, pumped thermal systems comprisingmultiple working fluid systems, or working fluid flow paths areprovided. In some cases, pumped thermal system components in the chargeand discharge modes may be the same. For example, the samecompressor/turbine pair may be used in charge and discharge cycles.Alternatively, one or more system components may differ between chargeand discharge modes. For example, separate compressor/turbine pairs maybe used in charge and discharge cycles. In one implementation, thesystem has one set of heat exchangers, and a common set of HTS and CTStanks which are charged or discharged by two pairs or sets ofcompressors and turbines. In another implementation, the system has acommon set of HTS and CTS tanks, but separate sets of heat exchangersand separate sets of compressors and turbines.

Pumped thermal systems with recuperation, utilization of externalsources of heat, cold and/or waste heat/cold may benefit from havingseparate compressor/turbine pairs as a result of operation ofturbomachinery over large and/or different temperature ranges in chargeand discharge modes. For example, temperature changes between charge anddischarge cycles may lead to a thermal adjustment period or otherdifficulties during transition between the cycles (e.g., issues orfactor related to metallurgy, thermal expansion, Reynolds number,temperature dependent compression ratios, tip clearance and/or bearingfriction etc.). In another example, turbomachinery (e.g., turbomachineryused in systems with recuperation) may operate over a relatively lowpressure ratio (e.g., with relatively few compression stages) but overrelatively large temperature during both compression and expansion. Thetemperature ranges may change (e.g., switch as in FIGS. 8A and 8B)between charge and discharge modes. In some cases, the operation overlarge temperature ranges during compression and/or expansion maycomplicate design of a combined compressor/turbine for both charge anddischarge. Furthermore, recuperation, waste heat/cold incorporationand/or other pumped thermal system features may reduce the compressionratio of the compressor/turbine in the charge cycle and/or the dischargecycle, thereby reducing the cost associated with duplicatingcompressor/turbine sets.

FIGS. 14A and 14B show pumped thermal systems with separate compressor1/turbine 3 pairs for charge mode C and discharge mode D. The separatecompressor/turbine pairs may or may not be ganged on a common mechanicalshaft. In this example, the compressor/turbine pairs C and D can haveseparate shafts 10. The shafts 10 may rotate at the same speed or atdifferent speeds. The separate compressor/turbine pairs or working fluidsystems may or may not share heat exchangers (e.g., the heat exchangers2 and 4).

In the example in FIG. 14A, the system has a common set of HTS tanks 6and 7 and CTS tanks 8 and 9. The system has separate pairs of heatexchangers 2 and 4 and separate compressor 1/turbine 3 pairs for thecharge mode C and the discharge mode D. The HTS and CTS storage mediaflow paths for the charging cycle are shown as solid black lines. TheHTS and CTS storage media flow paths for the discharge cycle are shownas the dashed grey lines.

In the example in FIG. 14B, the system, shown in a charge configuration,has one set of heat exchangers 2 and 4, and a common set of HTS tanks 6and 7 and CTS tanks 8 and 9. The HTS and CTS tanks can be charged by acompressor/turbine set C, or discharged by a compressor/turbine set D,each set comprising a compressor 1 and a turbine 3. The system mayswitch between the sets C and D using valves 83. In the example in FIG.14A, the system, again shown in a charge configuration, has a common setof HTS tanks 6 and 7 and CTS tanks 8 and 9. The HTS and CTS tanks can becharged by the charge set C that includes a first set of the heatexchangers 2 and 4, the compressor 1 and the turbine 3. The HTS and CTStanks can be discharged by switching to a separate discharge set C thatincludes a second set of the heat exchangers 2 and 4, the compressor 1and the turbine 3.

In one example, if the charge and discharge sets of compressors andturbines in FIGS. 14A and 14B are not operated at the same time, thecharge and discharge sets may share a common set of heat exchangers thatare switched between the turbomachinery pairs using the valves 83. Inanother example, if the charge and discharge turbomachinery sets orpairs in FIGS. 14A and 14B are operated at the same time (e.g., in orderfor one set to charge, following intermittent generation, and the otherset to discharge at the same time, following load), then each set ofturbomachinery may have a dedicated set of heat exchangers. In thisinstance, the charge and discharge sets may or may not share a set ofHTS and CTS tanks.

In some implementations, separate compressor/turbine sets or pairs mayadvantageously be used in pumped thermal systems used with intermittentand/or variable electric power inputs. For example, a firstcompressor/turbine set can be used in a charge cycle that follows windand/or solar power (e.g., electric power input from wind and/or solarpower systems) while a second compressor/turbine set can be used in adischarge cycle that follows load (e.g., electric power output to apower grid). In this configuration, pumped thermal systems placedbetween a power generation system and a load may aid in smoothingvariations/fluctuations in input and/or output power requirements.

Hybrid Pumped Thermal Systems

In accordance with another aspect of the disclosure, pumped thermalsystems can be augmented by additional energy conversion processesand/or be directly utilized as energy conversion systems without energystorage (i.e., as power generation systems). In some examples, pumpedthermal systems herein can be modified to allow for direct powergeneration using natural gas, Diesel fuel, petroleum gas (e.g.,propane/butane), dimethyl ether, fuel oil, wood chips, landfill gas orany other combustible substance (e.g., fossil fuel or biomass) foradding heat to the working fluid on a hot side of a working fluid cycle,and a cold side heat sink (e.g., water) for removing heat from theworking fluid on a cold side of the working fluid cycle.

FIGS. 15A and 15B show pumped thermal systems configured in generationmode. In some examples, pumped thermal systems herein can be modified byadding two additional heat exchangers 40 and 41, four additional valves19 a, 19 b, 19 c and 19 d, a heat sink (e.g., a water cooling system;water from a fresh water reservoir such as a river, a lake or areservoir; salt water from a salt water reservoir such as a sea or anocean; air cooling using radiators, fans/blowers, convection; or anenvironmental heat sink such as ground/soil, cold air etc.) 42, and aheat source (e.g., a combustion chamber with a fuel-oxidant mixture) 43.The heat source 43 can exchange heat with a first of the two additionalheat exchangers 40, and the heat sink 42 can exchange heat with a secondof the two additional heat exchangers 41. The heat source 43 may be usedto for exchanging heat with the working fluid 20.

The heat source 43 may be a combustion heat source. In some examples,the combustion heat source can comprise a combustion chamber forcombusting a combustible substance (e.g., a fossil fuel, a syntheticfuel, municipal solid waste (MSW) or biomass). In some cases, thecombustion chamber may be separate from the heat exchanger 40. In somecases, the heat exchanger 40 may comprise the combustion chamber. Theheat source 43 may be a waste heat source, such as, for example wasteheat from a power plant, an industrial process (e.g., furnace exhaust).

In some examples, a solar heater, a combustion heat source, a waste heatsource, or any combination thereof may be used for heating the hot sideheat storage fluid and/or the working fluid. In an example, the workingfluid can be heated directly using any of these heat sources. In anotherexample, the hot side heat storage fluid (or HTS medium) can be heatedusing any of these heat sources. In another example, the hot side heatstorage fluid (or HTS medium) can be heated in parallel with the workingfluid using any of these heat sources.

The pumped thermal systems in FIGS. 15A and 15B may be operated ashybrid systems. For example, the valves 19 a, 19 b, 19 c and 19 d can beused to switch between two modes. When the valves are in a firstposition, the system can operate as a pumped thermal storage system(e.g., closed system in charge/discharge mode). In this configuration,the working fluid 20 (e.g., argon or air) can exchange heat with an HTSmedium (e.g., molten salt) in the hot side heat exchanger 2 and with aCTS medium (e.g., hexane) in the cold side heat exchanger 4. When thevalves are in a second position, the system can operate as a powergeneration system (e.g., open system in generation mode). In thisconfiguration, the heat exchangers 2 and 4 may be bypassed, and theworking fluid 20 can exchange heat with the combustion chamber 43 in thehot side heat exchanger 40 and with the heat sink 42 in the cold sideheat exchanger 41. Any description of configuration and/or design ofheat transfer processes (e.g., heat transfer in heat exchangers)described herein in relation to pumped thermal systems may also beapplied to hybrid pumped thermal systems, and vice versa. For example,the heat sink 42, the heat source 43, the heat exchangers 40 and 41,and/or the quantity of heat transferred on the cold side and/or the hotside may be configured to decrease or minimize entropy generationassociated with heat transfer processes and/or to maximize systemefficiency.

In some implementations, the hybrid systems may operate in storage andgeneration modes simultaneously. For example, the valves 19 a, 19 b, 19c and 19 d can be configured to allow a given split between a workingfluid flow rate to the heat exchangers 40 and 41 and a working fluidflow rate to the heat exchangers 2 and 4. Alternatively, the hybridsystems may operate exclusively in storage mode, or exclusively ingeneration mode (e.g., as a natural gas peaking plant). In some cases,the split between modes may be selected based on system efficiency,available electric power input (e.g., based on availability), desiredelectric power output (e.g., based on load demand) etc. For example,thermal efficiency of an ideal system (i.e., assuming isentropiccompression and expansion processes, ideal heat transfer processes)operating exclusively in generation mode can be the thermal efficiencyof a working fluid undergoing an ideal Brayton cycle. In some examples,thermal efficiencies in hybrid systems of the disclosure (e.g.,exclusive and/or split mode operation) can be at least about 10%, atleast about 20%, at least about 30%, at least about 40%, at least about50%, at least about 60%, or more.

The heat source 43 may be used for exchanging heat with an HTS medium(e.g., a molten salt). For example, the combustion heat source 43 may beused for heating the HTS medium 21. In some instances, rather than usingthe combustion heat source 43 for exchanging heat in the heat exchanger40 or for directly exchanging heat between flue gases from thecombustion heat source and the working fluid, the combustion heat source43 may be used to heat up the HTS medium 21 between the two HTS tanks 7and 6.

FIG. 15C is a schematic flow diagram of hot side recharging in a pumpedheat cycle through heating by the heat source 43 (e.g., combustion heatsource, waste heat source). In an example, the heat source 43 is a wasteheat source, such as a waste heat source from a refinery or otherprocessing plant. In an example, the heat source 43 is obtained fromcombusting natural gas in order to ensure the delivery of electricityeven if the pumped thermal system runs out of charged storage media. Forexample, recharging of the hot side storage media using the heat source43 may provide an advantage over recharging using electricity or othermeans (e.g., the price of electricity at the time may be too high). Theheat source 43 can be used to heat up the HTS medium 21 from thetemperature T₁ in the tank 7 to the temperature T₁ ⁺ in the tank 6. TheHTS medium can then be used in the CFX 2 for exchanging heat with theworking fluid to in a discharge cycle, such as, for example, thedischarge cycles in FIGS. 9G and 9H.

In some examples, such as, for example, when the CTS medium is acombustible substance such as a fossil fuel (e.g., hexane or heptanes),burning of the CTS medium stored in the CTS tanks (e.g., the tanks 8 and9) may be used for providing thermal energy for heating the HTS mediumas shown, for example, in FIG. 15C or for operation of the cycles in theconfigurations shown, for example, in FIGS. 15A and 15B.

The systems of the disclosure may be capable of operating both in anelectricity only storage cycle (comprising heat transfer with an HTSmedium and a CTS medium below ambient temperature) and in a heat engineto ambient cycle, where, in a discharge mode, heat is input from the HTSmedium to the working fluid and rejected to the ambient environmentrather than to the CTS medium. This capability may enable the use ofheating of the HTS with combustible substances (e.g., as shown in FIG.15C) or the use of solar heating of the HTS (e.g., as shown in FIG. 9C).Heat rejection to ambient may be implemented using, for example, thedischarge cycles in FIGS. 9G and 9H. In some cases, heat may be rejectedto the environment with the aid of the ITS medium 61 or the intermediatemedium 62.

The systems of the disclosure may be capable of operating both in anelectricity only storage cycle (comprising heat transfer with an HTSmedium and a CTS medium below ambient temperature) and in a refrigeratorto ambient cycle, where, in a cold side recharge mode, heat is takenfrom the CTS medium and rejected to the ambient environment in order tolower the temperature of the CTS. This capability may be paired with theuse of solar heating of the HTS (e.g., as shown in FIG. 9C). Heatrejection to ambient may be implemented using, for example, the coldside recharge cycle in FIG. 9F.

Aspects of the disclosure may be synergistically combined. For example,the systems capable of operating both in an electricity only storagecycle and in a heat engine to ambient cycle and/or the systems capableof operating both in an electricity only storage cycle and in arefrigerator to ambient cycle may comprise a recuperator. Anydescription in relation to such hybrid systems without a recuperator mayreadily be applied to hybrid systems with a recuperator at least in someconfigurations. In some instances, the hybrid systems may be implementedusing, for example, the parallel, valved configuration in FIG. 15A. Forexample, in such configurations, the counter-flow heat exchangers 2 inFIGS. 9A, 9B and 9F may be implemented as separate counter-flow heatexchangers 67 for exchanging heat with the ambient environment, and maybe used in combination with hot side counter-flow heat exchangers 2 ofthe disclosure. Similarly, the counter-flow heat exchangers 4 in FIGS.9G and 9H may be implemented as separate counter-flow heat exchangers 67for exchanging heat with the ambient environment, and may be used incombination with cold side counter-flow heat exchangers 4 of thedisclosure.

In some implementations, the systems herein may be configured to enableswitching between different cycles of the disclosure using a shared setof valves and pipes. For example, the system may be configured to switchbetween the electricity only charge cycle (such as shown in, forexample, FIG. 7C), the electricity only discharge cycle (such as shownin, for example, FIG. 7D), the cold side recharging refrigeration cycle(such as shown in, for example, FIG. 9F), and the heat engine to ambientcycle (such as shown in FIG. 9H).

Schematic diagrams of exemplary piping and valve configurations forachieving working fluid flow in various counter-flow thermodynamiccycles are shown in FIGS. 13A, 13B, 13C and 13D. Such configurations maybe advantageous as they may reuse at least a portion, or a substantialportion or majority of the same piping for the working fluid between asubset of modes or between all modes implemented in a given system.While the working fluid changes may change direction between cycles, thetemperature profile of the heat exchangers 2 (e.g., hot side CFX), 5(e.g., recuperator) and 4 (e.g., cold side CFX) can be kept constant,partially constant, or substantially or fully constant by changing thedirection in which the HTS medium 68 and the CTS medium 69 are pumped asneeded, and by matching the heat fluxes of the working fluid, the HTSmedium 68 and the CTS medium 69 appropriately. For example, the cyclesin FIGS. 13A, 13B, 13C and 13D can share a common set of pipes 66 andvalves 65 that the working fluid can flow through in all four cycles. Insome cases, implementation of a given cycle (e.g., which of the cyclesis run) may depend on what position the valves 65 are in and/or thedirection and flow rate of the HTS medium 68, the CTS medium 69 and anambient temperature cooling fluid 70 (e.g., the intermediate fluid 57,the ITS medium 61 or the intermediate fluid 62) through the heatexchangers.

The symbols 65 in FIGS. 13A-13D represent either three-way or four-wayvalves. The white quadrants indicate the direction in which the workingfluid is allowed to pass, black quadrants indicate where flow isblocked. For example, a valve symbol with white quadrants on left andright and black quadrants on top and bottom will allow the working fluidto pass from left to right and not in any other direction. The paths 66represent a series of pipes in which the working fluid 20 flows. Thepipes 66 may in some cases be thermally insulated and/or rated at agiven pressure or range of pressures. The pipes 66 may be so designed asto minimize losses due to turbulence and or pressure drop. Working fluidflow is indicated by light arrows inside the pipes 66.

FIG. 13A is a schematic diagram of an exemplary piping and valveconfiguration for achieving the electricity only charge cycle such asshown in, for example, FIG. 7C. FIG. 13B is a schematic diagram of anexemplary piping and valve configuration for achieving the electricityonly discharge cycle such as shown in, for example, FIG. 7D. FIG. 13C isa schematic diagram of an exemplary piping and valve configuration forachieving the cold side recharging refrigeration cycle such as shown in,for example, FIG. 9F. This configuration may also apply to the cold siderecharging refrigeration cycle in FIG. 9A (e.g., a refrigeration toambient cycle). In an example, the working fluid in FIG. 13C does notflow or pass through the recuperator 5. In some examples, an additionalset of valves and pipes may be used to implement a cold side rechargingrefrigeration cycle where the working fluid passes through therecuperator 5. FIG. 13D is a schematic diagram of an exemplary pipingand valve configuration for achieving heat engine to ambient cycle suchas shown in, for example, FIG. 9H. This configuration may also apply tothe heat engine to ambient cycle in FIG. 9G. In an example, the workingfluid in FIG. 13D flows or passes through the recuperator 5. In someexamples, an additional set of valves and pipes may be used to implementa heat engine to ambient cycle where the working fluid does not passthrough the recuperator 5.

The examples in FIGS. 13A, 13B, 13C and 13D illustrate an operatingprinciple of a valving and piping system configured such that thedifferent cycles of the system can be interchanged while maintaining thesame, partially the same, or substantially the same temperature profileacross at least one of the counter-flow heat exchangers, across a subsetof the counter-flow heat exchangers, or across all of the counter-flowheat exchangers in the system. This may be advantageous as the heatcapacities of the heat exchangers themselves can mean that changingtheir temperature profiles between cycles may result in additionalsystem losses. The example schematics also demonstrate the designprinciple of striving to keep the total amount of piping and valving toa minimum by reusing them for different cycles where possible. Invarious examples, different configurations of piping and valves, and/ora different number of pipes or valves may be used.

Pumped Thermal Systems with Pressure Regulation Power Control

In an aspect of the disclosure, the pressure of working fluids in pumpedthermal systems can be controlled to achieve power control. In anexample, the power provided to a closed system in charge mode and/or thepower extracted from the closed system in discharge and/or generationmode (e.g., work input/output using the shaft 10) is proportional to themolar or mass flow rate of the circulating working fluid. The mass flowrate is proportional to density, area, and flow velocity. The flowvelocity can be kept fixed in order to achieve a fixed shaft speed(e.g., 3600 rpm in accordance with power grid requirements). Thus, asthe pressure of the working fluid changes, the mass flow rate and thepower can change. In an example, as the mass flow rate increases in adischarge and/or generation mode, more load should be added to thesystem to maintain a constant speed of the rotating shaft, and viceversa. In another example, if load is reduced during operation in adischarge and/or generation mode, the reduced load can cause the shaftspeed to increase, thus increasing the mass flow rate. For some periodof time, before the heat stored in the thermal capacity of the heatexchangers themselves is dissipated, this increased mass flow rate canlead to an increase in the power delivered, in turn increasing the shaftspeed. The shaft speed and the power can continue to increaseuncontrollably, resulting in a runaway of the rotating shaft. In someexamples, pressure regulation may enable control, and thus stabilizationof runaway, through adjustment of the amount (e.g., density) ofcirculating working fluid in accordance with system requirements. In anexample where shaft speed (and turbomachinery, such as a turbine,attached to the shaft) begins to run away, a controller can reduce themass of circulating working fluid (e.g., mass flow rate) in order todecrease the power delivered, in turn decreasing the shaft speed.Pressure regulation may also allow for an increase in mass flow rate inresponse to an increase in load. In each of these instances, the flowrates of the HTS and CTS media through the heat exchangers can bematched to the heat capacity of the working fluid passing through theheat exchangers.

In some examples, the working fluid pressure in the closed system can bevaried by using an auxiliary working fluid tank in fluid communicationwith the closed system. In this configuration, power input/output can bedecreased by transferring the working fluid from the closed cycle loopto the tank, and power input/output can be increased by transferring theworking fluid from the tank to the closed cycle loop. In an example,when the pressure of the working fluid is decreased, less heat can betransferred between the thermal storage tanks on the hot and cold sidesof the system as a result of the decreased mass flow rate and less powercan be input to/output by the system.

As the pressure of the working fluid is varied, the compression ratiosof turbomachinery components may remain substantially unchanged. In somecases, one or more operating parameters and/or configuration (e.g.,variable stators, shaft speed) of turbomachinery components can beadjusted in response to a change in working fluid pressure (e.g., toachieve a desired performance of the system). Alternatively, one or morepressure ratios may change in response to a change in working fluidpressure.

In some cases, reduced cost and/or reduced parasitic energy consumptionmay be achieved using the power control configuration relative to otherconfigurations (e.g., using a choke valve for controlling the flow ofthe working fluid). In some examples, variation of working fluidpressure while keeping the temperature and flow velocity constant (ornear-constant) may lead to negligible entropy generation. In someexamples, an increase or decrease in system pressure may lead to changesin, for example, turbomachinery efficiencies.

FIG. 16 shows an example of a pumped thermal system with power control.The temperature of the working fluid on the hot and cold sides of thesystem may remain constant or near-constant for a given period of timeregardless of working fluid mass flow rate due to large heat capacitiesof the heat exchangers 2 and 4 and/or the hot and cold side thermalstorage media in the tanks 6, 7, 8 and 9. In some examples, the flowrates of the HTS and CTS media through the heat exchangers 2 and 4 arevaried in concert with a change in the pressure of the working fluid inorder to keep the temperatures in the heat exchangers and working fluidoptimized over longer time periods. Thus, pressure can be used to varythe mass flow rate in the system. One or more auxiliary tanks 44 filledwith the working fluid 20 (e.g., argon or argon-helium mix) can be influid communication with a hot (e.g., high pressure) side of the pumpedthermal system and/or a cold (e.g., low pressure) side of the pumpedthermal system. In some examples, the auxiliary tank can be in fluidcommunication with the working fluid adjacent to an inlet of thecompressor 1 and/or adjacent to an outlet of the compressor 1. In someexamples, the auxiliary tank can be in fluid communication with theworking fluid adjacent to an inlet of the turbine 3 and/or adjacent toan outlet of the turbine 3. In further examples, the auxiliary tank canbe in fluid communication with the working fluid in one or morelocations system (e.g., one or more locations on the high pressure sideof the system, on the low pressure side of the system, or anycombination thereof). For example, the auxiliary tank can be in fluidcommunication with the working fluid on a high pressure side and a lowpressure side of the closed cycle. In some cases, the fluidcommunication on the high pressure side can be provided after thecompressor and before the turbine. In some cases, the fluidcommunication on the low pressure side can be provided after the turbineand before the compressor. In some instances, the auxiliary tank cancontain working fluid at a pressure intermediate to the high and lowpressures of the system. The working fluid in the auxiliary tank can beused to increase or decrease the amount of working fluid 20 circulatingin the closed cycle of the pumped thermal system. The amount of workingfluid circulating in the closed cycle loop can be decreased by bleedingthe working fluid from the high pressure side of the closed cycle loopinto the tank through a fluid path containing a valve or mass flowcontroller 46, thereby charging the tank 44. The amount of working fluidcirculating in the closed cycle loop can be increased by bleeding theworking fluid from the tank into the low pressure side of the closedcycle loop through a fluid path containing a valve or mass flowcontroller 45, thereby discharging the tank 44.

Power control over longer timescales may be implemented by changing thepressure of the working fluid and by adjusting the flow rates of the hotside 21 and cold side 22 thermal storage fluids through the heatexchangers 2 and 4, respectively.

In some examples, flow rates of the thermal storage media 21 and/or 22may be controlled (e.g., by a controller) to maintain given heatexchanger inlet and outlet temperatures. In some examples, a firstcontroller(s) may be provided for controlling the flow rates (e.g., massflow rates) of thermal storage media, and a second controller may beprovided for controlling the mass flow rate (e.g., by controlling mass,mass flow rate, pressure etc.) of the working fluid.

Pumped Thermal Systems with Pressure-Encased Motor/Generator

In another aspect of the disclosure, pumped thermal systems with apressure-encased motor/generator are provided. The pressure-encasedmotor/generator may be provided as an alternative to configurationswhere a shaft (also “crankshaft” herein) penetrates through a workingfluid containment wall (where it can be exposed to one or morerelatively high pressure differentials) in order to connect to amotor/generator outside the working fluid containment wall. In somecases, the shaft may be exposed to pressures and temperatures of theworking fluid in the low pressure portion of the working fluid cycle, inthe high pressure portion of the working fluid cycle, or both. In somecases, crankshaft seal(s) capable of holding back the pressures whichthe crankshaft is exposed to inside the working fluid containment wallcan be difficult to manufacture and/or difficult to maintain. In somecases, a rotating seal between high and low pressure environments may bedifficult to achieve. Thus, coupling of the compressor and turbine tothe motor/generator can be challenging. In some implementations, themotor/generator can therefore be placed entirely within the low pressureportion of the working fluid cycle, such that the exterior pressurevessel or working fluid containment wall may not need to be penetrated.

FIG. 17 shows an example of a pumped thermal system with a pressureencased generator 11. The motor/generator is encased within the pressurevessel or working fluid containment wall (shown as dashed lines) andonly feed-through electric leads 49 penetrate through the pressurevessel. A thermal insulation wall 48 is added between themotor/generator 11 and the working fluid in the low pressure portion ofthe cycle. The technical requirements for achieving an adequate sealthrough the thermal insulation wall can be less stringent due to thepressure being the same on both sides of the thermal insulation wall(e.g., both sides of the thermal insulation wall can be located in thelow pressure portion of the cycle). In an example, the low pressurevalue can be about 10 atm. In some cases, the motor/generator may beadapted for operation at elevated surrounding pressures. An additionalthermal insulation wall 50 can be used to create a seal between theoutlet of the compressor 1 and the inlet of the turbine 3 in the highpressure portion of the cycle. In some examples, placing themotor/generator on the cold side of the pumped thermal systems may bebeneficial to the operation of the motor/generator (e.g., cooling of asuperconducting generator).

Pumped Thermal Systems with Variable Stator Pressure Ratio Control

A further aspect of the disclosure relates to control of pressure inworking fluid cycles of pumped thermal systems by using variablestators. In some examples, use of variable stators in turbomachinerycomponents can allow pressure ratios in working fluid cycles to bevaried. The variable compression ratio can be accomplished by havingmovable stators in the turbomachinery.

In some cases, pumped thermal systems (e.g., the systems in FIGS. 8A and8B) can operate at the same compression ratio in both the charge and thedischarge cycles. In this configuration, heat can be rejected (e.g., tothe environment) in section 38 in the charge cycle and in section 39 inthe discharge cycle, wherein the heat in section 38 can be transferredat a lower temperature than the heat in section 39. In alternativeconfigurations, the compression ratio can be varied when switchingbetween the charge cycle and the discharge cycle. In an example,variable stators can be added to both the compressor and the turbine,thus allowing the compression ratio to be tuned. The ability to varycompression ratio between charge and discharge modes may enable heat tobe rejected at the lower temperature only (e.g., heat may be rejected insection 38 in the charge cycle but not in section 39 in the dischargecycle). In some examples, a greater portion (or all) of the heatrejected to the environment is transferred at a lower temperature, whichmay increase the roundtrip efficiency of the system.

FIG. 18 is an example of variable stators in a compressor/turbine pair.The compressor 1 and the turbine 3 can both have variable stators, sothat the compression ratio for each can be tuned. Such tuning mayincrease roundtrip efficiency.

The compressor and/or the turbine can (each) include one or morecompression stages. For example, the compressor and/or the turbine canhave multiple rows of repeating features distributed along itscircumference. Each compression stage can comprise one or more rows offeatures. The rows may be arranged in a given order. In one example, thecompressor 1 and the turbine 3 each comprise a sequence of a pluralityof inlet guide vanes 51, a first plurality of rotors 52, a plurality ofstators 53, a second plurality of rotors 52 and a plurality of outletguide vanes 54. Each plurality of features can be arranged in a rowalong the circumference of the compressor/turbine. The configuration(e.g., direction or angle) of the stators 53 can be varied, as indicatedin FIG. 18.

The compressor/turbine pair can be matched. In some cases, an outletpressure of the compressor can be about the same as an inlet pressure ofthe turbine, and an inlet pressure of the compressor can be about thesame as the outlet pressure of the turbine; thus, the pressure ratioacross the turbine can be the same as the pressure ratio across thecompressor. In some cases, the inlet/outlet pressures and/or thepressure ratios may differ by a given amount (e.g., to account forpressure drop in the system). The use of variable stators on both thecompressor and the turbine can allow the compressor and the turbine toremain matched as the compression ratio is varied. For example, usingthe variable stators, operation of the compressor and the turbine canremain within suitable operating conditions (e.g. within a given rangeor at a given point on their respective operating maps) as thecompression ratio is varied. Operation within given ranges or at givenpoints on turbomachinery operating maps may allow turbomachineryefficiencies (e.g., isentropic efficiencies) and resulting roundtripstorage efficiency to be maintained within a desired range. In someimplementations, the use of variable stators can be combined with othermethods for varying the compression ratios (e.g. variable shaft rotationspeed, bypassing of turbomachinery stages, gears, power electronics,etc.).

Pumped Thermal System Units Comprising Pumped Thermal System Subunits

A further aspect of the disclosure relates to control of charging anddischarging rate over a full range from maximum charging/power input tomaximum discharging/power output by building composite pumped thermalsystem units comprised of pumped thermal system subunits. In someexamples, pumped thermal systems may have a minimum power input and/oroutput (e.g., minimum power input and/or minimum power output) above 0%of maximum power input and/or output (e.g., maximum power input and/ormaximum power output), respectively. In such cases, a single unit byitself may be able to continuously ramp from the minimum power input tothe maximum power input and from the minimum power output to the maximumpower output, but may not be able to continuously ramp from the minimumpower input to the minimum power output (i.e., from the minimum powerinput to zero power input/output, and from zero power input/output tothe minimum power output). An ability to continuously ramp from theminimum power input to the minimum power output may enable the system tocontinuously ramp from the maximum power input to the maximum poweroutput. For example, if both the output power and the input power may beturned down all the way to zero during operation, the system may be ableto continuously vary the power consumed or supplied across a range fromthe maximum input (e.g., acting as a load on the grid) to the maximumoutput (e.g., acting as a generator on the grid). Such functionality mayincrease (e.g., more than double) the continuously rampable range of thepumped thermal system. Increasing the continuously rampable range of thepumped thermal system may be advantageous, for example, whencontinuously rampable power range is used as a metric for determiningthe value of grid assets. Further, such functionality may enable thesystems of the disclosure to follow variable load, variable generation,intermittent generation, or any combination thereof.

In some implementations, composite pumped thermal system units comprisedof multiple pumped thermal system subunits may be used. In some cases,each subunit may have a minimum power input and/or output above 0%. Thecontinuous ramping of the power from the maximum power input to themaximum power output may include combining a given quantity of thesubunits. For example, a suitable (e.g., sufficiently large) number ofsubunits may be needed to achieve continuous ramping. In some examples,the number of subunits can be at least about 2, 5, 10, 20, 30, 40, 50,100, 200, 500, 750, 1000, and the like. In some examples, the number ofsubunits is 2, 5, 10, 20, 30, 40, 50, 100, 150, 200, 250, 300, 350, 400,450, 500, 550, 600, 650, 700, 750, 800, 850, 900, 950, 1000 or more.Each subunit may have a given power capacity. For example, each subunitcan have a power capacity that is less than about 0.1%, less than about0.5%, less than about 1%, less than about 5%, less than about 10%, lessthan about 25%, less than about 50%, or less than about 90% of the totalpower capacity of the composite pumped thermal system. In some cases,different subunits may have different power capacities. In someexamples, a subunit has a power capacity of about 10 kW, 100 kW, 500 kW,1 MW, 2 MW, 5 MW, 10 MW, 20 MW, 50 MW, 100 MW, or more. The continuousramping of the power from the maximum power input to the maximum poweroutput may include controlling each subunit's power input and/or output(e.g., power input and/or power Output) separately. In some cases, thesubunits may be operated in opposing directions (e.g., one or moresubunits may operate in power input mode while one or more subunits mayoperate in power output mode). In one example, if each pumped thermalsystem subunit can be continuously ramped between a maximum power inputand/or output down to about 50% of the maximum power input and/oroutput, respectively, three or more such pumped thermal system subunitsmay be combined into a composite pumped thermal system unit that can becontinuously ramped from the maximum input power to the maximum outputpower. In some implementations, the composite pumped thermal system maynot have a fully continuous range between the maximum input power andthe maximum output power, but may have an increased number of operatingpoints in this range compared to a non-composite system.

Energy Storage System Units Comprising Energy Storage System Subunits

A further aspect of the disclosure relates to control of charging anddischarging rate over a full range from maximum charging/power input tomaximum discharging power output by building composite energy storagesystem units comprised of energy storage system subunits. In someexamples, energy storage systems may have a minimum power input and/oroutput (e.g., minimum power input and/or minimum power output) above 0%of maximum power input and/or output (e.g., maximum power input and/ormaximum power output), respectively. In such cases, a single unit byitself may be able to continuously ramp from the minimum power input tothe maximum power input and from the minimum power output to the maximumpower output, but may not be able to continuously ramp from the minimumpower input to the minimum power output (i.e., from the minimum powerinput to zero power input/output, and from zero power input/output tothe minimum power output). An ability to continuously ramp from theminimum power input to the minimum power output may enable the system tocontinuously ramp from the maximum power input to the maximum poweroutput. For example, if both the output power and the input power may beturned down all the way to zero during operation, the system may be ableto continuously vary the power consumed or supplied across a range fromthe maximum input (e.g., acting as a load on the grid) to the maximumoutput (e.g., acting as a generator on the grid). Such functionality mayincrease (e.g., more than double) the continuously rampable range of theenergy storage system. Increasing the continuously rampable range of theenergy storage system may be advantageous, for example, whencontinuously rampable power range is used as a metric for determiningthe value of grid assets. Further, such functionality may enable thesystems of the disclosure to follow variable load, variable generation,intermittent generation, or any combination thereof.

In some implementations, composite energy storage system units comprisedof multiple energy storage system subunits may be used. In someexamples, any energy storage system having power input/outputcharacteristics that may benefit from a composite configuration may beused. In some examples, systems having power input and/or power outputcharacteristics that may benefit from a composite configuration mayinclude various power storage and/or generation systems such as, forexample, natural gas or combined cycle power plants, fuel cell systems,battery systems, compressed air energy storage systems, pumpedhydroelectric systems, etc. In some cases, each subunit may have aminimum power input and/or output above 0%. The continuous ramping ofthe power from the maximum power input to the maximum power output mayinclude combining a given quantity of the subunits. For example, asuitable (e.g., sufficiently large) number of subunits may be needed toachieve continuous ramping. In some examples, the number of subunits canbe at least about 2, 5, 10, 20, 30, 40, 50, 100, 200, 500, 750, 1000,and the like. In some examples, the number of subunits is 2, 5, 10, 20,30, 40, 50, 100, 150, 200, 250, 300, 350, 400, 450, 500, 550, 600, 650,700, 750, 800, 850, 900, 950, 1000 or more. Each subunit may have agiven power capacity. For example, each subunit can have a powercapacity that is less than about 0.1%, less than about 0.5%, less thanabout 1%, less than about 5%, less than about 10%, less than about 25%,less than about 50%, or less than about 90% of the total power capacityof the composite energy storage system. In some cases, differentsubunits may have different power capacities. In some examples, asubunit has a power capacity of about 10 kW, 100 kW, 500 kW, 1 MW, 2 MW,5 MW, 10 MW, 20 MW, 50 MW, 100 MW, or more. The continuous ramping ofthe power from the maximum power input to the maximum power output mayinclude controlling each subunit's power input and/or output (e.g.,power input and/or power output) separately. In some cases, the subunitsmay be operated in opposing directions (e.g., one or more subunits mayoperate in power input mode while one or more subunits may operate inpower output mode). In one example, if each energy storage systemsubunit can be continuously ramped between a maximum power input and/oroutput don to about 50% of the maximum power input and/or output,respectively, three or more such energy storage system subunits may becombined into a composite energy storage system unit that can becontinuously ramped from the maximum input power to the maximum outputpower. In some implementations, the composite energy storage system maynot have a fully continuous range between the maximum input power andthe maximum output power, but may have an increased number of operatingpoints in this range compared to a non-composite system.

Control Systems

The present disclosure provides computer control systems (orcontrollers) that are programmed to implement methods of the disclosure.FIG. 19 shows a computer system 1901 (or controller) that is programmedor otherwise configured to regulate various process parameters of energystorage and/or retrieval systems disclosed herein. Such processparameters can include temperatures, flow rates, pressures and entropychanges.

The computer system 1901 includes a central processing unit (CPU, also“processor” and “computer processor” herein) 1905, which can be a singlecore or multi core processor, or a plurality of processors for parallelprocessing. The computer system 1901 also includes memory or memorylocation 1910 (e.g., random-access memory, read-only memory, flashmemory), electronic storage unit 1915 (e.g., hard disk), communicationinterface 1920 (e.g., network adapter) for communicating with one ormore other systems, and peripheral devices 1925, such as cache, othermemory, data storage and/or electronic display adapters. The memory1910, storage unit 1915, interface 1920 and peripheral devices 1925 arein communication with the CPU 1905 through a communication bus (solidlines), such as a motherboard. The storage unit 1915 can be a datastorage unit (or data repository) for storing data. The computer system1901 can be operatively coupled to a computer network (“network”) 1930with the aid of the communication interface 1920. The network 1930 canbe the Internet, an internet and/or extranet, or an intranet and/orextranet that is in communication with the Internet. The network 1930 insome cases is a telecommunication and/or data network. The network 1930can include one or more computer servers, which can enable distributedcomputing, such as cloud computing. The network 1930, in some cases withthe aid of the computer system 1901, can implement a peer-to-peernetwork, which may enable devices coupled to the computer system 1901 tobehave as a client or a server.

The computer system 1901 is coupled to an energy storage and/orretrieval system 1935, which can be as described above or elsewhereherein. The computer system 1901 can be coupled to various unitoperations of the system 1935, such as flow regulators (e.g., valves),temperature sensors, pressure sensors, compressor(s), turbine(s),electrical switches, and photovoltaic modules. The system 1901 can bedirectly coupled to, or be a part of, the system 1935, or be incommunication with the system 1935 through the network 1930.

The CPU 1905 can execute a sequence of machine-readable instructions,which can be embodied in a program or software. The instructions may bestored in a memory location, such as the memory 1910. Examples ofoperations performed by the CPU 1905 can include fetch, decode, execute,and writeback.

With continue reference to FIG. 19, the storage unit 1915 can storefiles, such as drivers, libraries and saved programs. The storage unit1915 can store programs generated by users and recorded sessions, aswell as output(s) associated with the programs. The storage unit 1915can store user data, e.g., user preferences and user programs. Thecomputer system 1901 in some cases can include one or more additionaldata storage units that are external to the computer system 1901, suchas located on a remote server that is in communication with the computersystem 1901 through an intranet or the Internet.

The computer system 1901 can communicate with one or more remotecomputer systems through the network 1930. For instance, the computersystem 1901 can communicate with a remote computer system of a user(e.g., operator). Examples of remote computer systems include personalcomputers (e.g., portable PC), slate or tablet PC's (e.g., Apple® iPad,Samsung® Galaxy Tab), telephones, Smart phones (e.g., Apple® iPhone,Android-enabled device, Blackberry®), or personal digital assistants.The user can access the computer system 1901 via the network 1930.

Methods as described herein can be implemented by way of machine (e.g.,computer processor) executable code stored on an electronic storagelocation of the computer system 1901, such as, for example, on thememory 1910 or electronic storage unit 1915. The machine executable ormachine readable code can be provided in the form of software. Duringuse, the code can be executed by the processor 1905. In some cases, thecode can be retrieved from the storage unit 1915 and stored on thememory 1910 for ready access by the processor 1905. In some situations,the electronic storage unit 1915 can be precluded, andmachine-executable instructions are stored on memory 1910.

The code can be pre-compiled and configured for use with a machine havea processer adapted to execute the code, or can be compiled duringruntime. The code can be supplied in a programming language that can beselected to enable the code to execute in a pre-compiled or as-compiledfashion.

Aspects of the systems and methods provided herein, such as the computersystem 1901, can be embodied in programming. Various aspects of thetechnology may be thought of as “products” or “articles of manufacture”typically in the form of machine (or processor) executable code and/orassociated data that is carried on or embodied in a type of machinereadable medium. Machine-executable code can be stored on an electronicstorage unit, such memory (e.g., read-only memory, random-access memory,flash memory) or a hard disk. “Storage” type media can include any orall of the tangible memory of the computers, processors or the like, orassociated modules thereof, such as various semiconductor memories, tapedrives, disk drives and the like, which may provide non-transitorystorage at any time for the software programming. All or portions of thesoftware may at times be communicated through the Internet or variousother telecommunication networks. Such communications, for example, mayenable loading of the software from one computer or processor intoanother, for example, from a management server or host computer into thecomputer platform of an application server. Thus, another type of mediathat may bear the software elements includes optical, electrical andelectromagnetic waves, such as used across physical interfaces betweenlocal devices, through wired and optical landline networks and overvarious air-links. The physical elements that carry such waves, such aswired or wireless links, optical links or the like, also may beconsidered as media bearing the software. As used herein, unlessrestricted to non-transitory, tangible “storage” media, terms such ascomputer or machine “readable medium” refer to any medium thatparticipates in providing instructions to a processor for execution.

Hence, a machine readable medium, such as computer-executable code, maytake many forms, including but not limited to, a tangible storagemedium, a carrier wave medium or physical transmission medium.Non-volatile storage media include, for example, optical or magneticdisks, such as any of the storage devices in any computer(s) or thelike, such as may be used to implement the databases, etc. shown in thedrawings. Volatile storage media include dynamic memory, such as mainmemory of such a computer platform. Tangible transmission media includecoaxial cables; copper wire and fiber optics, including the wires thatcomprise a bus within a computer system. Carrier-wave transmission mediamay take the form of electric or electromagnetic signals, or acoustic orlight waves such as those generated during radio frequency (RF) andinfrared (IR) data communications. Common forms of computer-readablemedia therefore include for example: a floppy disk, a flexible disk,hard disk, magnetic tape, any other magnetic medium, a CD-ROM, DVD orDVD-ROM, any other optical medium, punch cards paper tape, any otherphysical storage medium with patterns of holes, a RAM, a ROM, a PROM andEPROM, a FLASH-EPROM, any other memory chip or cartridge, a carrier wavetransporting data or instructions, cables or links transporting such acarrier wave, or any other medium from which a computer may readprogramming code and/or data. Many of these forms of computer readablemedia may be involved in carrying one or more sequences of one or moreinstructions to a processor for execution.

It is to be understood that the terminology used herein is used for thepurpose of describing specific embodiments, and is not intended to limitthe scope of the present invention. It should be noted that as usedherein, the singular forms of “a”, “an” and “the” include pluralreferences unless the context clearly dictates otherwise. In addition,unless defined otherwise, all technical and scientific terms used hereinhave the same meaning as commonly understood by one of ordinary skill inthe art to which this invention belongs.

While preferable embodiments of the present invention have been shownand described herein, it will be obvious to those skilled in the artthat such embodiments are provided by way of example only. Numerousvariations, changes, and substitutions will now occur to those skilledin the art without departing from the invention. It should be understoodthat various alternatives to the embodiments of the invention describedherein may be employed in practicing the invention. It is intended thatthe following claims define the scope of the invention and that methodsand structures within the scope of these claims and their equivalents becovered thereby.

What is claimed is:
 1. A method comprising: operating a pumped thermal system in a charging cycle at a first compression ratio, wherein the pumped thermal system comprises a working fluid circulating through, in sequence, a compressor system, a hot side heat exchanger, a turbine system, and a cold side heat exchanger, wherein the working fluid is in thermal contact with a hot thermal storage (“HTS”) medium in the hot side heat exchanger and the working fluid is in thermal contact with a cold thermal storage (“CTS”) medium in the cold side heat exchanger; and operating the pumped thermal system in a discharging cycle at a second compression ratio different than the first compression ratio.
 2. The method of claim 1, wherein the compressor system comprises at least a first compressor and the second compressor.
 3. The method of claim 2, wherein operating the pumped thermal system in the charging cycle at the first compression ratio comprises circulating the working fluid through the first compressor and the second compressor in parallel, and wherein operating the pumped thermal system in the discharging cycle at the second compression ratio comprises circulating the working fluid through the first compressor and the second compressor in series.
 4. The method of claim 2, wherein operating the pumped thermal system in the charging cycle at the first compression ratio comprises circulating the working fluid through the first compressor and not the second compressor, and wherein operating the pumped thermal system in the discharging cycle at the second compression ratio comprises circulating the working fluid through the second compressor and not the first compressor.
 5. The method of claim 2, wherein operating the pumped thermal system in the charging cycle at the first compression ratio comprises circulating the working fluid through the first compressor and not the second compressor, and wherein operating the pumped thermal system in the discharging cycle at the second compression ratio comprises circulating the working fluid through the first compressor and the second compressor.
 6. The method of claim 1, wherein the turbine system comprises at least a first turbine and a second turbine.
 7. The method of claim 6, wherein operating the pumped thermal system in the charging cycle at the first compression ratio comprises circulating the working fluid through the first turbine and the second turbine in series, and wherein operating the pumped thermal system in the discharging cycle at the second compression ratio comprises circulating the working fluid through the first turbine and the second turbine in series.
 8. The method of claim 6, wherein operating the pumped thermal system in the charging cycle at the first compression ratio comprises circulating the working fluid through the first turbine and not the second turbine, and wherein operating the pumped thermal system in the discharging cycle at the second compression ratio comprises circulating the working fluid through the second turbine and not the first turbine.
 9. The method of claim 6, wherein operating the pumped thermal system in the charging cycle at the first compression ratio comprises circulating the working fluid through the first turbine and the second turbine, and wherein operating the pumped thermal system in the discharging cycle at the second compression ratio comprises circulating the working fluid through the first turbine and not the first turbine.
 10. The method of claim 1, wherein the compressor system comprises a compressor, wherein operating the pumped thermal system in the charging cycle at the first compression ratio comprises operating the compressor at a first rotation speed, and wherein operating the pumped thermal system in the discharging cycle at the second compression ratio comprises operating the compressor at a second rotation speed different than the first rotation speed.
 11. The method of claim 1, wherein the turbine system comprises a turbine, wherein operating the pumped thermal system in the charging cycle at the first compression ratio comprises operating the turbine at a first rotation speed, and wherein operating the pumped thermal system in the discharging cycle at the second compression ratio comprises operating the turbine at a second rotation speed different than the first rotation speed.
 12. The method of claim 1, wherein the compressor system comprises at least a first compressor and a second compressor, wherein the turbine system comprises at least a first turbine and a second turbine, wherein operating the pumped thermal system in the charging cycle at the first compression ratio comprises circulating the working fluid through the first compressor and the first turbine and not circulating the working fluid through the second compressor and the second turbine, and wherein operating the pumped thermal system in the discharging cycle at the second compression ratio comprises circulating the working fluid through the second compressor and the second turbine and not circulating the working fluid through the first compressor and the first turbine.
 13. The method of claim 1, wherein the compressor system comprises a compressor comprising a variable pressure stator, wherein operating the pumped thermal system in the charging cycle at the first compression ratio comprises tuning the variable stator to operate the compressor at a third compression ratio across the compressor, and wherein operating the pumped thermal system in the discharging cycle at the second compression ratio comprises tuning the variable stator to operate the compressor at a fourth compression ratio across the compressor.
 14. The method of claim 1, wherein the turbine system comprises a turbine comprising a variable pressure stator, wherein operating the pumped thermal system in the charging cycle at the first compression ratio comprises tuning the variable stator to operate the turbine at a third compression ratio across the turbine, and wherein operating the pumped thermal system in the discharging cycle at the second compression ratio comprises tuning the variable stator to operate the turbine at a fourth compression ratio across the turbine.
 15. The method of claim 1, wherein the first compression ratio is a value such that $\frac{T_{1}^{+}}{T_{1}} = {\psi_{C}^{1/\eta_{cp}}.}$
 16. The method of claim 1, wherein the second compression ratio is a value such that $\frac{T_{1}^{+}}{T_{1}} = {\psi_{D}^{\eta_{tp}}.}$
 17. The method of claim 1, wherein the second compression ratio is greater than the first compression ratio.
 18. The method of claim 17, wherein the first compression ratio is less than 3 and the second compression ratio is greater than
 3. [0129]
 19. The method of claim 1, wherein the first compression ratio varies based at least in part on a temperature of the working fluid.
 20. The method of claim 1, wherein the second compression ratio varies based at least in part on a temperature of the working fluid. 